Active vehicle suspension

ABSTRACT

A method of on-demand energy delivery to an active suspension system is disclosed. The suspension system includes an actuator body, a hydraulic pump, an electric motor, a plurality of sensors, an energy storage facility, and a controller. The method includes disposing an active suspension system in a vehicle between a wheel mount and a vehicle body, detecting a wheel event requiring control of the active suspension; and sourcing energy from the energy storage facility and delivering it to the electric motor in response to the wheel event.

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a continuation of U.S. patent application Ser. No.15/432,901, filed Feb. 14, 2017, which is a continuation of U.S. patentapplication Ser. No. 14/859,892, filed Sep. 21, 2015, which is acontinuation application of U.S. patent application Ser. No. 14/213,860,filed Mar. 14, 2014, which claims the benefit of priority under 35U.S.C. § 119(e) of U.S. provisional application Ser. No. 61/789,600,filed Mar. 15, 2013, the contents of each of which are hereinincorporated by reference in their entirety.

BACKGROUND Field

The methods and systems described herein relate to improvements inactive vehicle suspension.

Art

Current active suspension systems can benefit from improvements inpower, efficiency, architecture, size, and compatibility, many of whichare described herein.

SUMMARY

In one embodiment, a method of on-demand energy delivery to an activesuspension system is disclosed. The suspension system includes anactuator body, a hydraulic pump, an electric motor, a plurality ofsensors, an energy storage facility, and a controller. The methodincludes disposing an active suspension system in a vehicle between awheel mount and a vehicle body, detecting a wheel event requiringcontrol of the active suspension; and sourcing energy from the energystorage facility and delivering it to the electric motor in response tothe wheel event.

It should be appreciated that the foregoing concepts, and additionalconcepts discussed below, may be arranged in any suitable combination,as the present disclosure is not limited in this respect. Further, otheradvantages and novel features of the present disclosure will becomeapparent from the following detailed description of various non-limitingembodiments when considered in conjunction with the accompanyingfigures.

BRIEF DESCRIPTION OF THE FIGURES

The accompanying drawings are not intended to be drawn to scale. In thedrawings, each identical or nearly identical component that isillustrated in various figures may be represented by a like numeral. Forpurposes of clarity, not every component may be labeled in everydrawing. In the drawings:

FIG. 1A depicts a graph of conventional semi-active suspensionforce/velocity range.

FIG. 1B depicts a graph of active suspension four-quadrant control.

FIG. 1C depicts a hydraulic active suspension actuator.

FIG. 1D depicts a graph of active suspension energy flow.

FIG. 1E depicts a graph of frequency-domain based rapid motor control ofan active suspension system.

FIG. 1F depicts suspension motor control on a per-event basis.

FIG. 1G depicts a block diagram of an electronic suspension system.

FIG. 1H depicts a Bode diagram of frequency versus magnitude of torquecommand correlation to body acceleration.

FIG. 2A depicts an active suspension control system block diagram.

FIG. 2B depicts voltage bus signaling and operating modes.

FIG. 2C depicts DC converter capacitor configurations.

FIG. 2D depicts active suspension system with processor-based controllerper wheel.

FIG. 2E depicts a processor-based controller per wheel of FIG. 2D.

FIG. 2F depicts a block diagram representing distributed vehicledynamics control.

FIG. 2G depicts a block diagram representing centralized vehicledynamics control.

FIG. 2H depicts a disconnect failure mode for an active suspensioncontrol system.

FIG. 2I depicts a short-circuit failure mode for an active suspensioncontrol system.

FIG. 2J depicts a 3-phase bridge circuit for controlling an activesuspension system.

FIG. 2K depicts an equivalent circuit of the circuit of FIG. 2J whenpower is removed.

FIG. 3A is an embodiment of a regenerative active/semi active damperthat comprises a side mounted hydraulic regenerative, active/semi activedamper valve with a rotary position sensor.

FIGS. 3B and 3C is an embodiment of a regenerative active/semi activeside mounted damper valve with a rotary position sensor.

FIG. 3D is an embodiment of a regenerative active/semi active damperthat comprises an inline mounted hydraulic regenerative, active/semiactive damper valve with a rotary position sensor.

FIG. 3E is an embodiment of a regenerative active/semi active inlinemounted damper valve with a rotary position sensor.

FIG. 4A is an embodiment of a monotube passive damper with a hydraulicbuffer.

FIG. 4B is an embodiment of a hydraulic buffer mounted in a piston headof a monotube passive damper.

FIG. 4C is an embodiment of a regenerative active/semi active damperwith a hydraulic buffer.

FIG. 4D is an embodiment of a hydraulic buffer mounted in a piston headof a regenerative active/semi active damper.

FIG. 5A is an embodiment of a regenerative active/semi active damperthat comprises a hydraulic regenerative, active/semi active damper valvein a monotube damper architecture with a passive throttle valve andpassive blow-off valve placed schematically in parallel and in serieswith the damper valve.

FIGS. 5B, 5C and 5D is an embodiment of a passive blow-off valve mountedin a piston head.

FIGS. 5E, 5F 5G, 5H, and 5I is an embodiment of a single acting throttlevalve mounted in the rebound chamber and compression chamber of aregenerative active/semi active damper.

FIGS. 5J, 5K and 5L is an embodiment of a dual acting throttle valvemounted in the rebound chamber of a regenerative active/semi activedamper.

FIGS. 5M, 5N, 5O, 5P, 5Q, 5R, 5S, and 5T is an embodiment of a divertervalve mounted in the rebound chamber and compression chamber of aregenerative active/semi active damper.

FIG. 5U is a curve of force/velocity of a regenerative active/semiactive damper without passive valve curve shaping.

FIG. 5V is a curve of force/velocity of a regenerative active/semiactive damper with passive throttle valve and passive blow-off valvecurve shaping.

FIG. 5W is a curve of force/velocity of a regenerative active/semiactive damper with passive diverter valve curve shaping.

FIG. 5X is an embodiment of a regenerative active/semi active damperthat comprises a hydraulic regenerative, active/semi active damper valvein a monotube damper architecture with a passive diverter valve placedschematically in parallel and in series with the damper valve.

FIG. 6A is an embodiment of a regenerative active/semi active damperthat comprises a hydraulic regenerative, active/semi active damper smartvalve with a top mounted board arrangement.

FIG. 6B is an embodiment of a regenerative active/semi active damperthat comprises a hydraulic regenerative, active/semi active damper smartvalve with a side mounted board arrangement.

FIG. 6C is an embodiment of a regenerative active/semi active smart witha top mounted board arrangement.

FIG. 6D is an embodiment of a regenerative active/semi active smartvalve with a side mounted board arrangement.

FIG. 6E demonstrates a generic Smart Valve architecture.

FIG. 6F demonstrates an integrated active suspension system thatpackages into the wheel-well of a vehicle, substantially between theupper 6-54 and lower 6-55 suspension members.

FIG. 6G demonstrates an embodiment of an inertia cancelling controlsystem for an active suspension.

FIG. 6H demonstrates an embodiment of the control signals of a torqueripple cancelling control system.

FIG. 7A shows a flow diagram for the process for self-calibration ofrotor position sensing.

FIG. 7B shows one possible implementation of the signal as used in anadaptive encoder model.

FIG. 7C explains the relationship between actual measured quantity (inthis case position) and the output of a typical sensor with errors ofthis type.

FIG. 7D shows a possible embodiment of the complete motor controlstrategy for an active suspension system.

FIG. 7E shows how the back EMF can be represented in DQ0 space by avector 7-30.

FIG. 7F depicts the shape of the filter represented by block 7-2 in FIG.7A.

FIG. 8A shows an embodiment of the system depicting “situational” activecontrol system, whereby the amount of active control used is dictateddirectly by the situation at hand.

FIG. 8B shows a possible way to describe the requirements and frequencyranges in a typical modern automobile, mapped against the frequency axis8-8.

FIG. 8C demonstrates subplots showing the distinction between a “hill”or a “bump” road feature of an active control system.

FIG. 8D depicts a scheme to allow the user to experience maximumcomfort, while also maintaining average output or regenerated power atthe expected target levels.

FIG. 8E shows an open loop control system approach for a “situational”active control system.

FIG. 8F shows an embodiment of the active control system where theinformation gathered from the front wheels is used to improve the eventdetection on the rear wheels.

FIG. 8G demonstrates a control system that improves controllability ofan adaptive suspension (either fully active or semi-active) by utilizinginformation from the front wheels in order to control the rear wheels.

DETAILED DESCRIPTION

On-Demand Energy Delivery to an Active Suspension

Disclosed is a method of on-demand energy delivery to an activesuspension system. The method includes disposing an active suspensionsystem in a vehicle between a wheel mount and a vehicle body, the activesuspension system including an actuator body, hydraulic pump, electricmotor, plurality of sensors, energy storage facility, and controller;detecting a wheel event requiring control of the active suspension; andsourcing energy from the energy storage facility and delivering it tothe electric motor in response to the wheel event. The energy in theactive suspension system is sourced from the energy storage facility anddelivered to the electric motor at a rate of 1 Hertz or faster. In anexample of the system, the actuator body is a multi-channel actuatorbody with one or more concentric tubes. In an example of the system, theenergy storage facility includes either a vehicle battery, supercapacitor, or hydraulic accumulator. In an embodiment, at least some ofthe energy in the system comes from the vehicle alternator. In anexample, the system generates electricity by transforming hydraulicfluid flow in the actuator due to the wheel event into electricity bydirecting the fluid flow to rotate the hydraulic pump which in turnrotates the electric motor for producing electric energy.

Also, disclosed is a control system for on-demand energy delivery to anactive suspension system. The system includes an active suspensionsystem including an actuator body, hydraulic pump, electric motor,plurality of sensors, energy storage facility, and motor controller. Thecontrol system dynamically updates the torque or speed setting in themotor controller in direct response to a wheel event.

Also disclosed is a hydraulic active suspension system for a vehicle.The hydraulic active suspension system includes a hydraulic actuator, ahydraulic pump in fluid communication with the hydraulic actuator, anelectric motor operatively coupled to the hydraulic pump, a controllerwhich can vary the kinematic characteristic of the electric motor, afluid circuit such that movement of the actuator in a first directionresults in rotary motion of the electric motor in a clockwise direction,and movement of the actuator in a second direction results in rotarymotion of the electric motor in a counterclockwise direction. Over halfof the fluid pumped by the hydraulic pump is used to move the piston,and not lost by bypassing the piston through one or more valves.

Also disclosed is a system that is able to push and pull a suspension.The system includes an electric motor, a hydraulic pump operativelyconnected to the electric motor, a hydraulic actuator operativelyconnected to the hydraulic pump, a controller that has one or moresensors for feedback. Motor control is updated to respond to wheel orvehicle body events in order to at least partially alter the hydraulicactuator force.

In an example, the motor torque is updated at a rate of or greater than1 Hertz. In an example, motor control is torque control of the motor. Inother examples, the motor control is velocity control of the motor orposition control of the motor. In an example, the hydraulic actuatorforce is partially controlled by one or more valves. In an example, oneor more valves is a pressure-operated valve, In alternate example, theone or more valves may be a inertia-operated valve; anacceleration-operated valve or an electronically controlled valve. In anexample, the electric motor can operate at least partially as agenerator and the hydraulic pump at least partially as a hydraulic motorin order to regenerate energy.

Controllable Valveless Active Suspension

Also disclosed is an active suspension system including an actuatorbody, hydraulic pump, electric motor, and plurality of sensors. Anelectric motor coupled to a hydraulic pump rotates, causing fluid toflow such that in a first direction it moves the actuator for reboundand in a second direction for compression; and a suspension controllerthat is adapted to control torque or speed of the electric motor inresponse to a sensed wheel event to one of mitigate and accentuate fluidflowing through the actuator, thereby effecting active suspension forceor position without requiring use of valves for disrupting the flow offluid through the actuator body. In an example of such active suspensionsystem, the actuator body is a multi-channel actuator body.

Energy Neutral Active Suspension Control System

Also disclosed is a method of energy neutral active suspension includingharvesting energy from suspension actuator movement; delivering theharvested energy to an energy source from which the suspension actuatorconditionally draws energy to effect active suspension; and consumingenergy from the energy source to control movement of the suspensionactuator for wheel events that result in actuator movement. Energyconsumption is regulated and limited so that harvested energysubstantially equals consumed energy over a time period that issubstantially longer than an average wheel event duration.

In an example of such system, the active control demand threshold is 75watts, such that net energy consumed or regenerated over a time periodthat is substantially longer than an average wheel event duration isless than or equal to 75 watts. In another example, the energy source iseither a vehicle electrical system, a lead acid vehicle battery, a supercapacitor, or a hydraulic actuator.

Self-Powered Adaptive Suspension

Also disclosed is a method of self-powered adaptive suspension includingharvesting energy from suspension actuator movement, storing theharvested energy in an energy storage facility from which the suspensionactuator conditionally draws energy to effect dynamic force changes inthe suspension, consuming energy from the energy storage facility tocontrol movement of the suspension actuator for wheel events that resultin actuator movement and adapting the dynamic force changes based on ameasure of availability of energy in the energy storage facility toensure that a level of the energy in the energy storage facility doesnot drop below a self-powered threshold.

Also disclosed is a self-powered adaptive suspension system including apiston disposed in a hydraulic housing, an energy recovery mechanismsuch that movement of the piston results in energy generation, a controlsystem that regulates force on the piston by varying an electricalcharacteristic of the energy recovery mechanism, an energy storagefacility to which harvested energy from the energy recovery mechanismstores energy and a control system that operates from energy stored inthe energy storage facility.

Active Truck Cab Stabilization

Also disclosed is an active suspension system for the suspended cabin ofa truck including a plurality of hydraulic actuators, each actuatorcapable of providing controlled force in three or more quadrants.Rotation of a hydraulic pump creates pressure in the hydraulicactuators.

Also disclosed is an active suspension system for the suspended cabin ofa truck including a plurality of hydraulic actuators, one or moresensors, each actuator including an electric motor operatively coupledto a hydraulic pump; and a controller that regulates speed or torque ofthe electric motor based on sensor inputs.

Using Voltage Bus Levels to Signal System Conditions

Also disclosed is a method including sensing a voltage level of a powerdistribution bus of an active vehicle suspension system, comparing thesensed voltage to a plurality of system condition voltage ranges todetermine a current system condition voltage range and determining asystem condition based on the determined current system conditionvoltage range. The system condition is one of net regeneration, bias lowenergy, fault handling and recovery, under voltage shutdown, and loaddump.

Disclosed is a method including sensing a voltage level of a powerdistribution bus of an active vehicle suspension system, comparing thesensed voltage to a plurality of system condition voltage ranges todetermine a current system condition voltage range and determining asystem condition based on the determined current system conditionvoltage range. The system condition biases energy usage command in theactive vehicle suspension system

DC Voltage Bus Level Defines [Vehicle] Energy System Capacity

Also disclosed is a method including sensing a voltage level of a powerdistribution bus on a vehicle including at least one energy storagedevice. The power distribution bus is connected to a vehicle primaryelectrical bus via a bi-directional voltage converter; and determiningan energy capacity of the power distribution bus by comparing the sensedvoltage level to a predefined voltage range from a minimum to a maximumvoltage. The energy capacity is determined to be high when the sensedvoltage level is substantially the same as the maximum voltage andlesser sensed voltages indicate lower energy capacity. In an example,the power distribution bus is for an active suspension system.

Voltage Failure Tolerant Smart Valve (Corner Controller Damper)

Also disclosed is an active vehicle suspension controller including aprocessor-base controller, a transistor-based motor controller; andcontrolling an active suspension including a hydraulic pump that isco-axial with an electric motor. The suspension controller operates theactive suspension as a substantially passive vehicle suspension damperwith non-zero damping force when power is not applied in a normallyoperating manner to the processor-based controller or to the actuator.

Also disclosed is an active vehicle suspension controller including aprocessor-based controller, a transistor-based motor controller; andcontrolling an active suspension including an electric motor. Thesuspension controller operates the active suspension as an adaptivesuspension (controllable damping in at least two quadrants) when poweris not applied in a normally operating manner to the processor-basedcontroller.

Super Capacitor Use in a Vehicle Active Suspension System

Also disclosed is an active vehicle suspension system including anelectric motor, a processor-based controller able to control a motortorque or speed, a suspension power bus at a higher voltage than thevehicle primary electrical bus, at least one DC-DC converter, asupercapacitor array including one or more supercapacitors configured inseries and/or parallel; and the supercapacitor array being connected toa DC-DC converter in the system such that the voltage across the arrayis less than the suspension power bus voltage.

Bi-Directional 48<->12 VDC Converter

Also disclosed is a power converter for vehicular applications includinga DC-DC bi-directional voltage converter disposed between a vehicleprimary electrical bus and a secondary power bus operated at a voltageequal to or greater than the vehicle primary electrical bus voltage; anda super capacitor disposed across the voltage converter. The voltageconverter facilitates charging the super capacitor via the secondarypower bus.

In an example, the current is dynamically regulated based on one or moresensor inputs or external command inputs. In an example, the sensorinput is a battery voltage level, or the external command input is adigital CAN bus signal.

Control Topology of an Active Suspension Including a Three-Phase ACController on a Local Power Bus (e.g. 48 VDC) which can be Coupled to aVehicle Power Bus

Also disclosed is an active vehicle suspension electrical systemincluding a plurality of corner controllers. Each corner controller maycontain a three-phase motor bridge that controls a brushless directcurrent (BLDC) motor, a DC-DC converter that maintains a suspensionpower bus voltage at less than 60 volts; and an energy storage deviceable to deliver electrical energy onto the suspension power bus.

Control Topology of an Active Suspension Including a Processor-BasedController Per Wheel

Also disclosed is an active vehicle suspension system including aplurality of corner controllers. Each corner controller controls asuspension actuator for a vehicle wheel. Control logic for thesuspension actuator is at least partially calculated in the cornercontroller. The plurality of corner controllers are connected via acontrol topology circuit and a gateway for connecting one or more of thecorner controllers with vehicle systems via the control topology circuitfor transferring vehicle dynamics information among the vehicle systemsand the plurality of corner controllers.

In an example, the control topology circuit is a CAN or FlexRay bus. Inan example, the vehicle dynamics information is transferred from thevehicle systems to the one or more corner controllers by means of a CANor FlexRay bus connection. In an example, the plurality of cornercontrollers communicate vehicle state information via the controltopology circuit. In an example, the vehicle state information is datafrom one or more sensors, or calculated vehicle state (e.g. wheel speedor roll rate).

Electric Motor Rotor Position Sensing in an Active Suspension

Disclosed is a hydraulic active or semi-active damper valve including anelectric motor operatively connected to a hydraulic motor and a rotaryposition sensor, whereby the rotary position signal is used to controlthe torque and speed of said electric motor.

In an example, the rotary position sensor includes a Hall effect sensorand source magnet. The axis of the sensor and source magnet are coaxialwith the rotational axis of the electric motor. Alternatively, the axesof the sensor and source magnet are off-axis with the rotational axis ofthe electric motor and the source magnet is of an annular construction.Further, the sensor is located in a sealed sensor body that isconstructed of a magnetic material that is held in rigid connection tothe valve body. Also, the sensor body may contain a sealed sensor shieldconstructed of a non-magnetic material. Also, there exists an air gapbetween the sensor shield and the sensor. The source magnet is locatedin close proximity to the sensor in a supported in a holder constructedof a non-magnetic material that is operatively connected to the motorrotor.

In an example, the rotary position sensor includes of an opticalposition sensor. In such example, the sensor is located in a sealedsensor body that is held in rigid connection to the valve body. Also,the sensor body may contain a sealed sensor shield constructed of anoptically clear material.

Magnetically Sensing Electric Motor Rotor Position Through a Diaphragm

Disclosed is a fluid-immersed rotor position sensing device including anelectric motor rotor that is immersed in a hydraulic fluid, a magnet onthe shaft of the rotor of the electric motor, a diaphragm constructed ofnon-magnetic material that establishes a dry region that does notcontain hydraulic fluid, and a wet region that may contain hydraulicfluid; and a magnetic encoder or Hall effect sensor that measuresmagnetic field in order to determine either relative or absoluteposition.

Sensing Rotor Position of a Fluid Immersed Electric Motor Shaft in anActive Suspension

Also disclosed is a method including disposing a magnetic sensor targeton a rotor of an electric motor that is immersed in a hydraulic fluid,disposing a diaphragm to establish a dry region that does not containhydraulic fluid so that the magnetic sensor target passes proximal tothe diaphragm at least once per revolution of the rotor, positioning amagnetic sensor in the dry region to facilitate detecting the magneticsensor target each time it passes proximal to the diaphragm, resultingin a series of detections that each represent a rotation of the rotor;and processing the series of detections with a filter and an algorithmthat calculates ratios of velocity to determine rotor position andacceleration.

Sensing Rotor Position of a Fluid Immersed Electric Motor Hydraulic PumpPower Pack

Disclosed is an electro-hydraulic power pack including a hydraulicpump/motor operatively connected to an electric motor/generator, whichis encased in the working fluid, and a rotary position sensor, wherebythe rotary position signal is used to control the torque and speed ofsaid electric motor/generator.

In an example, the rotary position sensor includes of a Hall effectsensor and source magnet.

In such example, the axis of the sensor and source magnet are coaxialwith the rotational axis of the electric motor. Alternatively, the axisof the sensor and source magnet are off-axis with the rotational axis ofthe electric motor and the source magnet is of an annular construction.Further, the sensor is located in a sealed sensor body that isconstructed of a magnetic material that is held in rigid connection tothe valve body. Also, the sensor body may contain a sealed sensor shieldconstructed of non-magnetic material. Also, there exists an air gapbetween the sensor shield and the sensor. The source magnet is locatedin close proximity to the sensor in a supported in a holder constructedof a non-magnetic material that is operatively connected to the motorrotor.

In an example, the rotary position sensor includes of an opticalposition sensor. The sensor is located in a sealed sensor body that isheld in rigid connection to the valve body. Further, the sensor body maycontain a sealed sensor shield constructed of an optically clearmaterial.

High Frequency Accumulator with an Active Suspension to Mitigate Effectof Low Energy High Frequency Events on an Active Suspension which canhave an in-Tube Accumulator

Disclosed is a method of mitigating low energy, high frequency events onan active suspension including disposing an accumulator in a fluid flowtube of a damper of the active suspension system, directing fluid intothe accumulator in response to an increase in pressure of the fluid,accumulating energy delivered by directing the fluid into theaccumulator to target maintaining a substantially constant fluidpressure in the fluid flow tube, damping said fluid and returning thefluid directed into the accumulator to the fluid flow tube in responseto a decrease in pressure of the fluid.

Also disclosed is a device for mitigating low energy, high frequencyevents in an active suspension, including a floating piston with a firstand second side, an accumulator volume that may include at least acompressible medium or mechanical spring in communication with firstside of piston, a fluid volume that is in communication with the secondside of the piston that may contain a mechanical spring, a fluid orificethat has a first side and second side. The first side of the fluidorifice is in communication with the fluid volume and the second side ofthe fluid orifice is in communication with the variable force chamber ofa damper.

In an example, the damper is passive or semi active. Further the damperis in combination with one of a monotube damper, a twin tube damper anda triple tube damper.

In an example, the damper is a hydraulic regenerative, active/semiactive damper. Further, the damper is in combination with one of amonotube damper, a twin tube damper and a triple tube damper.

Separate Rebound and Compression Throttle Valves

Disclosed is an active suspension system including a hydraulic motor andseparate rebound and compression throttle valves at the inlets of thehydraulic motor such that each valve closes fluid flow into thehydraulic motor at a predetermined flow rate.

Communication Via Pressure Between Throttle Valve and Blowoff Valve

Disclosed is an active suspension system configured with a throttlevalve and a blow-off valve. The throttle valve closes at a given flowrate, and the blow-off valve opens when pressure in the system increasesdue to the throttle valve closing (i.e. they are in communication viapressure changes in a hydraulic fluid between the valves).

Combination of Throttle Valve and Blow-Off Valve in One Diverter Valvein a Shock Absorber

Disclosed is a passive valve in an active suspension including a sealingwasher, a throttle body that may contain flow passages that is incommunication with the first side of a hydraulic damper piston to afirst port of a hydraulic pump, a seal body that may contain a floworifice that is in communication with a second side of a damper and asecond port of a hydraulic pump and a mechanical spring that places apreload on the sealing washer, so that the sealing washer normallycloses off the flow orifices in the seal body, while generating a flowrestriction from the first side of the piston to the flow passages ofthe throttle body.

In an example the damper is in combination with one of a monotubedamper, a twin tube damper and a triple tube damper.

Use of Leakage in a Hydraulic Motor to Facilitate Fluid By-Pass of theMotor and to Mitigate the Effect of Inertia

Disclosed is a method of inertia mitigation in an active suspensionsystem including configuring a hydraulic fluid flow path that includes ahydraulic motor. The hydraulic motor is adapted to allow leakage of thefluid without impacting rotation of the motor.

Shaping of Force/Velocity Curves of an Active Suspension Using PassiveValving

Disclosed is a method including shaping force/velocity response curvesof an active suspension system using passive valving.

Disclosed is a hydraulic active or semi-active damper that uses one ormore passive hydraulic valves in parallel to a hydraulic pump in orderto manipulate the force/velocity response of the system, such that theone or more valves are open at predetermined pressures.

In an example, the damper is of a monotube architecture. In an alternateexample, the damper is of a twin tube/triple tube architecture

Smart Valve

Disclosed is an active suspension (actuator) that includes of anelectric motor, an electronic controller that controls torque or speedin the electric motor; and one or more sensors. The electric motor,electronic controller, and one or more sensors are integrated into asingle actuator body.

Disclosed is an active suspension (actuator) that includes of anelectric motor, a hydraulic pump; and an electronic controller thatcontrols torque or speed in the electric motor. The electric motor,hydraulic pump, and electronic controller are integrated into a singleactuator body.

Disclosed is an integrated valve that includes of an electric motor; ahydraulic pump; and an electronic controller that controls torque orspeed in the electric motor. The electric motor, hydraulic pump, andelectronic controller are contained in a single housing. The singlehoused is a fluid-filled housing.

Disclosed is an integrated valve that includes of an electric motor, ahydraulic pump, an electronic controller that controls torque or speedin the electric motor, and one or more sensors. Rotation of the electricmotor drives rotation of the hydraulic pump.

Disclosed is an active suspension (actuator) that includes of anelectric motor, a hydraulic pump, and a hydraulic actuator with a pistondisposed in it. Fluid is communicated between the hydraulic actuator andthe hydraulic pump through the body of the hydraulic actuator.

Disclosed is an active suspension system for a vehicle that includes ofa hydraulic motor that produces variable flow or variable pressure ateach corner of the vehicle and an electric motor at each corner of thevehicle. The electric motor is controlled to directly control wheelmovement.

Disclosed is an active suspension (actuator) that includes of anactuator body including a piston rod, an electric motor, an electroniccontroller that controls torque or speed in the electric motor; and anactuator that may contain at least one passive valve that operates inparallel and/or series with the hydraulic motor. The electric motor andelectronic controller are packaged with the actuator body so as to fitwithin a vehicle wheel well.

Disclosed is an active suspension (actuator) in which a smart valve isintegrated with the actuator, and occupies a volume and shape that canfit within the wheel well and the damper top and bottom mounts.

Disclosed is an active suspension (actuator) in which the smart valve isintegrated with the actuator, and occupies a volume and shape that suchthat during full range of motion and articulation of the damper,adequate clearance is maintained between the smart valve and allsurrounding components.

Disclosed is an active suspension (actuator) in which the smart valve isco-axial with the damper body and connects to the damper top mount.

Disclosed is an active suspension (actuator) in which the smart valve isco-axial with the damper body and occupies a diameter substantiallysimilar to that of an automotive damper top mount and spring perch.

Disclosed is an active suspension pump and motor that is less than 8inches in diameter and 8 inches in depth.

Disclosed is a control system for reducing the effect of inertia in anactive suspension, in which positive feedback is used to add the desiredtorque command on an electric motor with the product of the angularacceleration and a number substantially close to the moment of inertiaof the rotating element coupled to the electric motor (including therotor mass, etc.).

Disclosed is a positive feedback control system loop that adjusts torqueof an electric motor in an active suspension in order to compensate forforce caused by acceleration of an inertial element.

Disclosed is a method for controlling an electric motor coupled to ahydraulic motor, in which a desired torque command is added to theinverse of a modeled signal of the resulting torque ripple (out of phasewith it) in order to cancel pressure ripple from the hydraulic motor.

Disclosed is an active suspension system in which an electric generatoris adapted to be controlled by a processor-based controller for applyingtorque to a shaft of a hydraulic motor applies data processingtechniques to detect noise patterns in a stream of data derived from amagnetic sensor that senses rotations of the electric generator rotor.The noise patterns are filtered out by selective position sensing.

Disclosed is an active suspension system in which an electric generatoris adapted to be controlled by a processor-based controller for applyingtorque to a shaft of a hydraulic motor performs real-time online nolatency calibration of a sensor based on off-line generated calibrationcurve for detecting a position of a rotor of the electric generator.

Disclosed is a method of producing high-accuracy calibration using a lowcost position sensor including capturing position detection data fromthe low cost position sensor; applying a filtering algorithm thatdetects and filters out noise patters in the position detection data;and generating ratios of filtered and unfiltered velocity measures fromthe position detection data to produce a measure of acceleration and arotational position of an electric generator rotor.

Disclosed is a method of establishing a rotational position coordinatesystem relative to a rotor-axis of a 3-phase brushless motor includingdetecting a position of a magnetic target disposed on a rotatable shaftof an electric generator rotor; applying a current through windings ofthe electric generator to produce a magnetic field; detecting a newposition of the magnetic target and mapping the applied current to thedetected position to establish the position coordinate system.

Disclosed is a smart valve for providing a suspension function for awheel of a vehicle including an electric motor, a hydraulic pump, anelectronic controller that controls torque or speed in the electricmotor. The electric motor, hydraulic pump, and electronic controller areintegrated into a single actuator body; and at least one algorithmexecutable by the electronic controller for handling vehicle dynamicsfrom the group consisting of: large event handling algorithms, wheelaction prediction based on leading wheel actions, mode-specific energymanagement algorithms, power/energy optimizing algorithms, vehicledynamics algorithms that accept power/energy as a variable, powerthrottling algorithms, power averaging algorithms, open-loop driverinput correction algorithms, feed-forward active suspension controlalgorithms based on a vehicle model, active suspension end-stop controlalgorithms, frequency dependent damping algorithms, automatic gainscontrol algorithms, and situational control algorithms.

On-Demand Energy Delivery to an Active Suspension

To deliver the full capabilities of active suspension includingoperating in all four quadrants of a vehicle suspension force-velocitygraph as depicted in FIG. 1B (e.g. rebound damping, compression damping,rebound pushing, and compression pulling), energy must be applied to thesuspension in response to a wheel event (i.e. movement of the wheelrelative to the vehicle or a force required by the suspension on thewheel that is not correlated with wheel motion, such as what is requiredduring handling maneuvers or changing loads). To approximate achieving adesired level of suspension performance, a system must ensure thatenergy needed to be applied is present or provided at an appropriatetime. One commonly known approach to ensuring energy is applied in atimely manner is to maintain an actively operating pump thatcontinuously pumps suspension fluid, and then using one or moreelectronically controlled valves to shuttle this fluid flow in order tomove the piston. However, such approaches require constantly consumingenergy (e.g. provided by the vehicle) to actively operate the pump,motor, or other suspension components. Even worse, many of these systemsmust bypass the piston using one or more valves when the active force isnot necessary, which oftentimes results in over half the fluid pumped bythe hydraulic pump being wasted in bypass.

In some advanced systems of the prior art, the speed of the pump may beadjusted every so often in response to changes in the general roadconditions or driver settings (e.g. sport mode or comfort mode setting).Several technical challenges have historically limited faster control ofthe pump including startup friction, rotational inertia, and limitationsin the electronic control system.

On-demand energy delivery to an active suspension system aims to achieveoutstanding suspension performance while consuming energy only whenneeded, such as in response to a wheel event. Suspension systems thatare capable of on-demand energy delivery may significantly reduce powerconsumption requirements over solutions that use a continuously poweredpump. Generally, on-demand energy delivery in active mode can beaccomplished by maintaining timely access to sufficient stored energy,or by exercising rapid control of torque in an active suspension'smotor, and on-demand energy deliver in a regeneration mode can beaccomplished by increasing energy recovery by increasing the dampingforce. By combining energy storage, such as with a vehicle battery orcapacitors, with dynamic energy generation from suspension action,on-demand energy can be achieved that also facilitates energy neutral ornearly energy neutral fully active suspension. An embodiment of asuspension system that is capable of dynamic energy generation fromsuspension system action is depicted in FIG. 1C as a hydraulic activesuspension actuator. In a fundamental expression of on-demand energydelivery via the system of FIG. 1C, fluid flow/pressure change inresponse to an emerging wheel event can be harvested and directed toaccelerate/decelerate a hydraulic motor that can both power and beresisted by an electric generator.

An example of on-demand energy delivery may be gleaned from FIG. 1D thatshows energy flow in active suspension. When an on-demand energydelivery-capable active suspension system experiences positive energyflow (when the graph is above the center line), a regenerationcapability may utilize this source of energy (such as during rebound) togenerate electricity. This may occur when fluid flowing past thehydraulic motor in FIG. 1C due to wheel rebound action is used to turnthe electric generator, thereby producing electricity that may be storedfor on-demand consumption, or instantaneously consumed. When anon-demand energy delivery capable active suspension system experiencesnegative energy flow (when the graph in FIG. 1D is below the centerline), the energy harvested during the positive flow cycle can beconsumed as needed (e.g. on-demand). Alternatively in on-demand energydelivery capable active suspension systems without regenerationcapability, energy can be consumed from a variety of source such asenergy storage devices or a vehicle's 12V or 48V electrical system. Thiscan be effected in the suspension actuator of FIG. 1D by applying acounter acting current into the generator, thereby resisting therotation of the hydraulic motor which in turn increases pressure in theactuator causing the wheel movement driving the demand to be mitigated.Alternatively, applying a current into the generator may cause theactuator to actively move in a desired direction. Also, energyconsumption might be required throughout a wheel event, such as when avehicle encounters a speed bump. Energy may be required to lift thewheel as it goes over a speed bump (that is, reduce distance betweenwheel and vehicle) and then push the wheel down as it comes off of thespeed bump to keep the vehicle more level throughout. However, reboundaction, such as the wheel returning to the road surface as it comes downoff of the speed bump may, fall into the positive energy flow cycle byharnessing the potential energy in the spring, thereby affording anopportunity to generate energy.

In order to improve the dynamics of such a system in some embodiments alow-inertia hydraulic pump such as a gerotor is used. In addition, theelectric motor coupled to the hydraulic pump may also be low inertia,such as by using an elongated but narrow diameter rotor of the motor(for example, where the diameter of the rotor is less than the height ofthe rotor). Additionally, the system may use features such as bearings,a low startup torque hydraulic pump, or hydrodynamic bearings in orderto reduce startup friction of the rotating assembly. Finally, anadvanced motor controller such as three-phase brushless DC motor bridgewith a fast control loop (1 kHz and above) may be used to rapidly updatemotor torque.

Controllable Valveless Active Suspension

The representative hydraulic active suspension actuator of FIG. 1C mayalso be valveless in the sense that no electronically controlledorifices are used. Rather than using active valves to mitigate orsuspend fluid flow in the actuator to counteract wheel/vehicle bodymovement in response to a wheel event, a valveless hydraulic activesuspension actuator may include sensing and control aspects thatfacilitate making adjustments in aspects of the actuator function with acontroller in response to sensed changes in the wheel (e.g. movement)and/or sensed changes in actuator fluid flow. In an example, a controlsystem may energize the electric generator to resist rotary motion ofthe hydraulic motor and therefore increase pressure in a portion of theactuator chamber to slow down or prevent movement of the wheel relativeto the vehicle body. This may be power consuming or power regenerating,depending on the configuration and force. Likewise, this hydraulicactive suspension can be controlled to induce movement of the wheelrelative to the body by causing the hydraulic motor to rotate, therebymoving fluid among the channels of the actuator chamber.

Energy Neutral Active Suspension Control System

By being aware of energy flow in active suspension, an example of whichis depicted in FIG. 1D, it is possible to extract and utilize (eitherthrough storage or consumption) at least a portion of energy produced bythe suspension while in a regeneration mode. This stored energy can thenbe available on-demand when a wheel event requires consumption. Storedenergy can be harvested and provided by, for example, an electronicsuspension system as depicted in FIG. 1G that incorporatesbi-directional energy transfer between a suspension system and a vehicleelectrical network as well as optional energy storage via a supercapacitor that spans the two electrical networks. The bidirectionalnature of such an electronic suspension system may effectively permitreturn of consumed energy to the vehicle electrical system thereby,causing the suspension system to be nearly energy neutral over time.

In an example of energy neutral active suspension control, energycaptured via regeneration from small amplitude and/or low frequencywheel events may be stored in the super capacitor of FIG. 1G. Once thesuper capacitor is fully charged, additional energy generated can eitherbe transferred to the vehicle power network (e.g. to charge the vehiclebattery) or merely .and dissipated as heat. When the suspension controlsystem requires energy, such as to resist movement of a wheel or toencourage movement of a wheel in response to a wheel event, energy maybe drawn from the super capacitor and from the vehicle power network viathe bidirectional power converter. Energy that is consumed to managevarious wheel events may be replaced through the charging functionalitydescribed above, effectively resulting in energy neutral activesuspension control. In another example of energy neutral activesuspension control, the amount of energy flow is measured over time andthe actuator forces are biased such that the total average consumedpower is less than or equal to +/−75 watts (consumed or regenerated).Such a control system is not limited to regenerative capable systems,and can be accomplished by biasing suspension forces in the semi-active“regenerative” zones as average consumed power approaches a numbersubstantially close to zero such as 75 watts.

The suspension system described herein whereby energy flow from thesuspension is stored and at a later time used to create force or motionin the suspension can also be realized with other means of energystorage, e.g. hydraulic accumulators or flywheels. In this embodiment,the energy never enters the electrical domain and is simply transferredfrom kinetic energy into potential energy stored through a mechanismenabling its gradual reconversion into kinetic energy at a preciseinstant in time and to a precise amount.

Self-Powered Adaptive Suspension

Through a combination of energy harvesting during regeneration, storage,dynamic generation, and on-demand consumption, a self-powered adaptivesuspension system is achievable. As shown in FIG. 1C, energy can beextracted from a suspension actuator by causing fluid in the actuator toflow past a hydraulic motor, thereby rapidly rotating the motor. Thehydraulic motor may be coupled, such as through a common shaft, to anelectric generator. As the hydraulic motor rotates a shaft in commonwith the electric generator, the generator may produce electricity thatcan be used directly and/or conditioned and stored for later consumptionby the suspension control system and/or to influence the rotation of thehydraulic motor, thereby causing the actuator to perform as component ofan active suspension. In this entire disclosure, the words hydraulicmotor and hydraulic pump are used interchangeably each to mean amechanism that either converts fluid flow into rotary motion, covertsrotary motion into fluid flow, or both. Similarly, the words electricgenerator and electric motor are used interchangeable each to mean amechanism that either converts electric current into rotary motion,coverts rotary motion into electric current, or both (e.g. the word“generator” may be used to indicate a device that is only capable ofproducing rotation from energy and not vice versa).

Such a self-powered suspension system may also automatically adapt howit controls the suspension elements to respond to a wheel event based onavailability of self-powered energy reserves and self-powered energygenerating capabilities. Adaptability of self-powered suspension mightbe beneficial in ensuring that as energy reserves begin to diminish,responses to some wheel events might transition from consuming energy toharvesting energy from the actuator movements. Likewise, as energyreserves diminish, suspension responsiveness might adjust to a moreenergy-conserving mode of operation until sufficient energy reserved canbe detected to resume “normal” active suspension operation.

In an example of self-powered adaptive suspension control, energycaptured via regeneration from small amplitude and/or low frequencywheel events may be stored in the super capacitor of FIG. 1G. When thesuspension control system requires energy, such as to resist movement ofa wheel at very low velocities substantially close to zero velocity, orto encourage movement of a wheel, in response to a wheel event, energymay be drawn from the super capacitor. Energy that is consumed to managevarious wheel events may be replaced through the charging functionalitydescribed above, effectively resulting in self-powered adaptivesuspension control.

In embodiments the self powered adaptive suspension only operates inquadrants one and three (the regenerative, damping quadrants), and inother embodiments the system is also able to enter the active quadrants(i.e. two and four) with an active force created by stored energy frompreviously captured regenerative energy.

Active Truck Cab Stabilization System

In some embodiments the vehicle suspension system may be the secondarysuspension of a truck cab. In these cases, the present inventive methodsand systems may be a fully active suspension system in order tostabilize the isolated cab of commercial trucks. Such a system mayinclude four active actuators such as those of FIG. 1C, with an actuatorat each corner of the cab. These actuators replace the passive dampersfound on most isolated truck cabs. In some trucks, the cab may beisolated on one side, with the other side hinged to the truck body.Here, the system might employ two active actuators to stabilize the cab.The system uses a plurality of sensors (e.g. accelerometers) and/orvehicle data (e.g. steering angle) in order to sense or predict cabmovement, and a control system sends commands to the actuators in orderto stabilize the cab. Such cab stabilization provides significantimprovement in comfort and may reduce maintenance requirements in thetruck.

Turning now to the figures, and initially FIG. 1C, which shows anembodiment of a four-quadrant fully active hydraulic actuator. A piston1-10 is disposed in a fluid-filled housing 1-9. Upon movement of thepiston, a piston head 1-11 forces fluid into and out of one or moreconcentric fluid flow tubes 1-12. These tubes allow fluid communicationbetween each of the two sides of the piston head and a hydraulic motor1-14. In the embodiment of FIG. 1C, the hydraulic motor may be agerotor, vane pump, internal or external gear pump, gerolor, hightorque/low speed gerotor motor, turbine pump, centrifugal pump, axialpiston pump, bent axis pump, or any other device that may act as ahydraulic pump or a hydraulic motor.

In the embodiment of FIG. 1C which uses a positive displacementhydraulic motor that may be back-driven, upon movement of the piston rod1-10 in a first direction, the piston head 1-11 pushes fluid into one ormore of the fluid flow tubes 1-12, which in turn communicates fluid toand spins a hydraulic motor 1-14.

Similarly, in a first mode rotation of the hydraulic motor 1-14 in afirst direction forces fluid through one or more of the fluid flow tubes1-12. This flow causes high pressure on a first side (the compressionvolume) of the piston head 1-11 and low pressure on a second side of thepiston head. This pressure differential applies a force on the pistonrod 1-10 in the extension direction. In a second mode, rotation of thehydraulic motor 1-14 in a second direction forces fluid through one ormore of the fluid flow tubes 1-12. This flow causes high pressure on thesecond side (the extension volume) of the piston head 1-112 and lowpressure on the first side of the piston head. This pressuredifferential applies a force on the piston rod 1-10 in the compressiondirection.

In the embodiment of FIG. 1C, the hydraulic motor 1-14 is operativelycoupled to an electric motor 1-15, which may be a BLDC motor such as athree-phase permanent magnet synchronous motor, a brushed DC motor, aninduction motor, dynamo, or any other type of device that convertselectricity into rotary motion and/or vice-versa. The coupling betweenthe electric motor and the hydraulic motor may be a simple shaft, or mayinclude one or more devices to alter the kinematic transfercharacteristic such as a clutch (velocity, electronically,directionally, or otherwise controlled), a shock-absorbing device suchas a spring pin, or a cushioning/damping device, however such devicesare not limited in this regard.

The operative coupling of the electric motor and hydraulic pump are suchthat applying energy to the terminals of the electric motor 1-15 mayresult in movement of the piston rod 1-10 if the resulting forcegenerated by the pressure created by the hydraulic pump (caused bytorque on the electric motor acting on the pump), and acting on thepiston head is sufficient to overcome the force on the piston rod. Insome embodiments, movement of the piston rod 1-10 also results inmovement of the electric motor 1-15, however, the present inventivemethods and systems are not limited in this regard. Additionally, insome embodiments secondary passive or electronic valving is includedwhich may in certain modes decouple piston movement from electric motormovement (in other words, movement of the piston head might not createan immediate and correlated movement of the electric motor).

Since fluid volume in the fluid-filled housing 1-9 changes as the piston1-10 enters and exits the housing, the embodiment of FIG. 1C includes anaccumulator 1-13 to accept the piston rod volume. In one embodiment,this accumulator is a nitrogen-filled chamber with a floating pistonable to move in the housing and sealed from the hydraulic fluid.

The embodiment of FIG. 1C may be adapted in order to accommodate anumber of fluid flow paths. In one embodiment, the fluid flow tubes 1-12may be pipes or hydraulic hoses. In another embodiment, the fluid flowtubes may be the concentric area between the inner and outer tubes of atwin-tube damper or the concentric area between each of the three tubesof a triple-tube damper. Twin tube and triple tube dampers are wellknown in the art. In other embodiments, fluid may flow such that thehydraulic motor always spins in a single direction. One way this may beaccomplished is by using one or more check valves, and/or one or moreelectronically controlled valves such as solenoid valves commonly usedin semi-active dampers.

In some embodiments similar to FIG. 1C it may be desirable that thehydraulic pump and piston movement may be completely decoupled incertain modes, or that the hydraulic pump rotate in a single direction.One such unidirectional pump system uses two electronically controlledvalves (such as solenoid-based proportional valves), three check valves,and a hydraulic pump operatively coupled to an electric motor. Thehydraulic pump moves fluid out of the output port and through a checkvalve that ensures fluid is not returned to the pump to back-drive it.This fluid is then in direct communication with both 1) an extensionvolume of the actuator body, and 2) one of the electronically controlledvalves. When open, this electronically controlled valve allows fluid topass through it to the compression volume of the actuator body. Thecompression volume is in selective fluid communication through a checkvalve to the extension volume. When the piston is compressed, fluidflows from the compression volume to the extension volume, but flowcannot go in the opposite direction. The compression volume is alsoconnected through a second electronically controlled valve to both theinput port of the hydraulic motor and an accumulator that is sized to beat least large enough to accommodate the piston rod volume introducedduring a compression stroke. In addition, there is a directional checkvalve that allows fluid to pass from the accumulator to the compressionvolume, but not vice versa. In this embodiment, the electronicallycontrolled valves may be opened and closed to create semi-active dampingand to directionally control the actuator's active force. The motor isconnected to an electronic control system that rapidly regulates motortorque to create appropriate pressure in the system on a per wheel eventbasis. This may be discussed in greater detail later in this disclosure.

In some embodiments an electric motor may be replaced by anengine-driven hydraulic motor. In these embodiments, it may be desirableto provide an electronically controlled clutch or a pressure bypass inorder to reduce engine load while high active actuator forces are notneeded. In a similar fashion to the rapid torque changes of the electricmotor, the hydraulic motor drive (either through an electronic clutch,an electronically-controlled hydraulic bypass valve, or otherwise), maybe rapidly controlled on a per wheel event basis in order to modulateenergy usage in the system.

FIG. 1A demonstrates a representative plot of the command authority 1-3of a semi-active suspension. That is, the system is able to create forcethat counteracts movement (a reactive force). The prior art disclosesseveral systems that are able to create such a range. In severalhydraulic systems of the prior art, a simple electronically controlledvalve is used to regulate fluid flow. In the closed state the dampingcurve is at full stiffness 1-2, and in the open state the damping curveis at full soft 1-1. Limitations on the high force due to leakage andlimitations on the low force due to fluid losses/friction limit theoperable range of such systems. Electronically controlled valvesolutions of the prior art consume energy to operate and dissipatedamping energy as heat.

In certain embodiments of the present inventive methods and systems,however, the system is able to create a force range 1-3 substantiallysimilar to that shown in FIG. 1A while generating instead of consumingenergy. One such system is described in FIG. 1C. If the terminals of theelectric motor 1-15 are left in an open circuit state (thus, with highimpedance between them), a damping curve similar to the full soft 1-1curve may be achieved. If the terminals of the electric motor see a lowimpedance, a damping curve similar to the full stiff 1-2 curve may beachieved. For damping curves between these bounds, a regenerativesemi-active system such as that described in FIG. 1C may generate energyfrom wheel movement. Description of the high and low impedance states isa functional description; in most embodiments this is achieved with aswitching power converter such as an H-bridge motor controller, wherethe switches are controlled to achieve the desired torquecharacteristic. This torque may be controlled in direct response to awheel event, thereby creating force only when necessary.

FIG. 1B demonstrates a representative plot of the command authority 1-8of a full active suspension such as the embodiment described in FIG. 1C.In particular, the device is able to create a force range 1-8 thatencompasses three or more quadrants of the force velocity chart. In thefirst quadrant 1-4, the system is able to create a reactive force torebound (extension) of the wheel of a vehicle. In the third quadrant 1-5the system is able to create a reactive force to compression of thewheel of a vehicle. As previously described, the embodiment of FIG. 1Cis able to generate energy in at least part of these quadrants. Thesystem is also able to create a force in at least one of the tworemaining quadrants: an active force to pull the wheel up 1-6 and/or anactive force to push the wheel down 1-7. In these quadrants, the systemmay consume energy (this includes electronic energy from the vehicle oran energy storage device such as a capacitor, or in the case of a systemutilizing hydraulic energy storage from an accumulator or similardevice, or in the case of a system utilizing other mechanical means ofenergy storage another device such as for example a flywheel). In someembodiments such a system has bidirectional energy flow; that is, inquadrants one and three energy is regenerated, and in quadrants two andfour energy is consumed.

Such a system capable of creating controlled force in at least three ofthe force/velocity quadrants may be deemed a full active suspension. Inat least one embodiment of FIG. 1C, the full active system is able tocreate force 1-8 in all four of the quadrants.

FIG. 1D shows power flow over time in an energy-neutral full activesuspension. For positive y-axis values 1-17, energy is regenerated. Fornegative y-axis values 1-18 energy is consumed. In some embodimentspower flow into or out of the full active suspension 1-16 is onlyregenerated (in the regeneration quadrant 1-17) while the force/velocitycommand of FIG. 1C is in the first 1-4 or third 1-5 quadrants(accounting for any delays from the mechanical and hydraulic system,including energy storage devices). However, in some embodiments powerflow into or out of the active suspension 1-16 may be consumed acrossall four quadrants of the force/velocity plot shown in FIG. 1CB.

FIG. 1D shows the results of an embodiment of a control system thatregulates the force of a full active suspension such that average poweris within a small window substantially close to zero (for example,within 75 watts of either regeneration or consumption over an extendedperiod of time). Such a control system may be considered an energyneutral control system.

The control system of an active suspension such as that shown in FIG. 1Cmay involve a variety of parameters such as wheel and body acceleration,steering input, braking input, and look-ahead sensors such as visioncameras, planar laser scanners, and the like. In one embodiment of anenergy neutral control system, the controller calculates a runningaverage of power (consumed or regenerated). The controller throttlesgains into the algorithm such that the system biases more towards theregenerative region 1-3 of the force/velocity chart in order to keep theaverage power (equal to the total energy, the integral of the powercurve 1-16 divided by time) within the neutral band 1-19 (herein alsoreferred to as the active control demand threshold). In some embodimentssuch a control system is used with a mechanical system Active controldemand threshold that has a regenerative region 1-3 that does notcapture electrical energy (it is either purely dissipative or energy iscaptured by means of a hydraulic accumulator). One such embodiment ofthe mechanical system involving two electronically controlled valves andthree check valves is disclosed above.

While embodiments described previously take an average of the singleactuator power flow, the present inventive methods and systems are notlimited in this regard. In some embodiments an average may be taken onthe sum of all actuators of the vehicle, or a subset of them.Additionally, the average may be over all time, between vehicle ignitionstarts, over a small time window, or over any other of a multitude oftime periods. In addition, the control system in some embodimentsincludes a safety mode where power limits are overridden duringavoidance, braking, fast steering, and when other safety-criticalmaneuvers are sensed.

A simple embodiment of a safety-critical maneuver detection algorithm isa trigger if the brake position is depressed beyond a certain thresholdBrake_pos_threshold and the derivative of the position (the brakedepression velocity) also exceeds a threshold Brake_vel_threshold. Aneven simpler system may utilize longitudinal acceleration thresholds.Another simple embodiment may utilize steering. In one application ofthis system, a fast control loop compares a threshold valueSteering_emergency_threshold to a factor derived by multiplying thesteering rate and a value from a lookup table indexed by the currentspeed of the vehicle. The lookup table may contain scalar values thatrelate maximum regular driving steering rate at each vehicle speed. Forexample, in a parking lot a quick turn is a conventional maneuver.However, at highway speeds the same quick turn input is likely a safetymaneuver where the suspension should disregard energy limits in order tokeep the vehicle stabilized. In another simple example, a vehiclerollover model for SUVs may be utilized that incorporates a number ofsensors such as lateral acceleration to change the suspension dynamicsif an imminent rollover condition is detected. In many real-worldapplications, a number of these heuristics (braking, steering,lane-departure/traffic detection sensors, deceleration, lateralacceleration, etc.) are fused together (such as by using fuzzy logic) tocome to a determination. The determination might not be binary, butrather, a scaling factor on the power limits.

FIG. 1E shows a frequency plot relating motor torque updates 1-23 withbody control 1-21 and wheel control 1-22 frequency bands. For a typicalpassenger vehicle, body movement occurs between zero and four Hertz,although higher-frequency body movement may occur well beyond this band.Wheel movement often occurs between eight and twenty Hertz, roughlycentered around 10 Hertz, although this differs from vehicle to vehicleand based on road conditions. A wheel event may be defined as any inputinto the wheel that causes a wheel and or body movement (including theresult of a steering input). From a frequency perspective, this oftenoccurs at roughly 0.5 Hertz and above, and may even occur at frequenciesin excess of one thousand Hertz. From a functional perspective, anyinstantaneous changes in commanded motor torque in response to a wheelevent (as measured by one or more sensors) may be considered a responseto a wheel event. To this regard the electro-hydraulic system is in someembodiments configured so as to achieve maximum stiffness between theelectric generator/motor and the hydraulic pump/motor. This has beenachieved by a using short, close-coupled shaft to connect the electricgenerator/motor to the hydraulic pump/motor. The connection of thehydraulic pump/motor to the shaft may also incorporate spring pinsand/or drive key features so as to reduce backlash between them. In theembodiment where the hydraulic pump/motor is a gerotor type, theassembly is configured so that the root and/or tip clearance can beeasily adjusted so as to reduce backlash and or leakage between theinner and outer gerotor elements. In the embodiment where a monotubedamper architecture is used, a high gas pre-charge (>35 bar) issometimes used, to increase the hydraulic fluid stiffness and hencereduce lag and latency. In other embodiments a gas pre-charge around 25bar may be used.

Referring again to the embodiment of FIG. 1C and using the control motortorque control system (1-23) referenced in FIG. 1E, the presentinventive methods and systems are able to update motor torque on a perwheel event basis. That is, the torque of the electric motor 1-15either 1) changes at an update rate greater than or equal to thefrequency at which wheel events occur, or 2) occurs in direct responseto a sensed event. With such rapid motor control, regulating torque onthe electric motor provides energy-efficient control.

In order to perform such rapid updates of motor torque, an electronicssystem capable of reacting to sensor inputs and dynamically changingmotor torque is required. One embodiment of such a system is disclosedin FIG. 1G. In this embodiment a full active suspension actuator 1-9 isconnected to a corner controller power electronics system 1-31. In theembodiment of FIG. 1G, the hydraulic motor/electric motor valve 1-30 maycontain a three-phase electric motor, and an encoder. The presentinventive methods and systems are not limited in this regard however,and may utilize a number of motor technologies, and sensorless controlinstead of an encoder. The corner controller 1-31 may include as inputsone or more sensors 1-32 such as wheel and body accelerometers, positionsensors, and vehicle CAN bus. The corner controller may further utilizeinformation derived from the electric motor (for example, calculatingactuator velocity by measuring electric motor velocity, calculatingactuator force by measuring electric motor current, and the like).Further sources of sensor data may include look-ahead sensors, and foractuators on the rear axle of the vehicle, information on the road fromthe front wheels may be used.

The corner controller 1-31 is able to control torque from the electricmotor of the full active suspension. One electrical topology embodimentof such a system is a three-phase bridge, with six MOSFET transistors.Each motor phase is connected to the junction between two MOSFETs inseries, with the high side MOSFET connected to the voltage rail and thelow side MOSFET connected to ground. Additionally, a controller rapidlypulse-width-modulates a control signal to the gate of each MOSFET inorder to drive the motor.

In the system-level embodiment of FIG. 1G, energy flows into and out ofthe corner controller on the suspension electrical bus 1-33, which maybe direct current. In some embodiments this voltage is held at a voltagehigher than that of the vehicle's electrical system, such as 48 volts or380 volts. In this configuration, smaller wires with lower current maybe used, which delivers a potential cost, weight, and integrationadvantage. In other embodiments this voltage is substantially similar tothe vehicle's electrical system voltage (12, 24 or 48 volts), which mayeliminate or downsize the need for a DC-DC converter. In certaincircumstances such as to reduce super capacitor cost, it may bedesirable to use a voltage lower than the vehicle's electrical system.

In some embodiments the suspension electrical bus 1-33 is voltageregulated, but the regulation allows bus voltage to fluctuate someamount from V_low to V_high. In the example of a nominal 48V busvoltage. V_low may be around 40V and V_high may be 50V.

While not shown in FIG. 1G, in one embodiment multiple actuators 1-9 andcorner controllers 1-31 share a common suspension electrical bus 1-33.In this way, if one actuator/corner controller pair is regeneratingenergy, another pair can be consuming this regenerated energy.

Modern vehicles are limited in their capacity to accept recuperativeelectrical energy from onboard devices, and to deliver large amounts ofenergy to onboard devices. In the former, regenerated energy may causevehicle electrical system voltage to rise higher than allowable, and inthe latter, large power draws may cause a voltage brownout, orunder-voltage condition for the vehicle. In order to deliver sufficientpower to the active suspension, or to capture a maximal amount ofregenerated energy, a form of energy storage may be used. Energy storagemay be in the form of batteries such as lithium ion batteries with acharge controller, ultra capacitors, or other forms of electrical energystorage. In the embodiment of FIG. 1G, supercapacitors 1-35 are used ina configuration where the negative terminal of the capacitor isconnected to a positive terminal of vehicle electrical system 1-36, andthe positive terminal is connected to the suspension electrical bus 1-33running at a voltage higher than the vehicle electrical system voltage.The present inventive methods and systems are not limited in thisregard, however, and the energy storage may be placed directly on thesuspension electrical bus or the vehicle electrical system.

In the embodiment of FIG. 1G, the suspension electrical bus 1-33interfaces with the vehicle's electrical system 1-36 and the vehicle'smain battery 1-37 through a bidirectional DC-DC converter 1-34. Severalbidirectional converters are known in the art and suitable for thistask, both galvantically isolated and non-galvantically isolatedtopologies. A few possible topologies include a synchronous buckconverter (where the freewheeling diode is replaced with a transistor),a transformer with fast-switching DC/AC converters on each side, andresonant converters, however the present inventive methods and systemsare not limited in this regard.

FIG. 1F demonstrates an active suspension motor torque control systemthat updates in response to wheel events. As can be seen in the chart,changes to the commanded motor torque 1-39 occur at a similar frequencyover time to body acceleration 1-38, which is caused by wheel eventssuch as bumps, hills, and potholes, and driver inputs such as turns,braking, etc.

FIG. 1H shows the same data in terms of frequency instead of time. Theshape of the motor torque magnitude command with respect to frequency1-41 roughly traces the shape of the body acceleration magnitude withrespect to frequency 1-40. This trace of the control algorithmdemonstrates that not only is commanded motor torque updated atfrequencies at least as high as wheel events are occurring, but alsothat there is high correlation between the motor torque magnitude andthe body acceleration magnitude.

In a simplified proportional control system embodiment, commanded motortorque 1-39 is updated at 10 Hz. At each update, commanded torque is setto be the current vertical body velocity (body acceleration put througha software integrator) multiplied by a scaling factor-k such that theactuator creates a force opposite to the body velocity. Such acontroller is optimized for improving body control.

In another simplified control system optimized for wheel control,commanded torque is set to be the current damper velocity (differentialmovement between the wheel and body) and multiplied by a factor-k inorder to counteract movement. Here, the system responds much like adamper. In some embodiments, body control and wheel control systems arefused together in order to provide full vehicle control.

In other embodiments commanded motor torque 1-39 is updated at slowerrates such as 0.5 Hz or faster rates such as 1 kHz. More complex controlsystems may also utilize other sensor data in addition or instead ofbody acceleration, and may include proportional, integral, derivative,and more complex feedback control schemes

Electronics

Using Voltage Bus Levels to Signal System Conditions

An active suspension system of a vehicle may be powered by a voltage busthat is controllably isolated from a primary vehicle voltage bus tofacilitate mitigating impact on the vehicle systems connected to theprimary voltage bus as the suspension system's demand for power can varysubstantially based on speed, road conditions, suspension performancegoals, and the like. As demand on the suspension voltage bus varies, thevoltage level of the bus may also vary, generally with the voltage levelincreasing when demand is low or in the case of recuperative systemswhen regeneration levels are high, and voltage decreasing when demand ishigh. By monitoring the voltage level of the suspension power bus, itmay be possible to determine, or at least approximate, suspension systemconditions. As noted above, a decreased voltage level on the suspensionsystem power bus may indicate a high demand for power to respond towheel events. This information may in turn allow a determination, orapproximation, of other information about the vehicle; for example, ahigh demand for power due to wheel events may in turn indicate that theroad surface is rough or sharply uneven, that the driver is engaging indriving behavior that tends to result in such wheel events, and thelike.

In embodiments, a suspension system power bus may transfer energy amongcorner controllers and a central converter as depicted in FIG. 2A. Eachcorner controller may independently monitor the suspension system powerbus to determine the overall system conditions for taking appropriateaction based on these system conditions, as well as monitoring any wheelevents being experienced locally for the wheel with which the cornercontroller is associated. A central converter connected to thesuspension system power bus may also connect to a vehicle primary powerbus that is driven by the vehicle battery. This connection may allow forenergy to be exchanged between the suspension system power bus and thevehicle primary power bus. In addition, the central converter and anyenergy storage either on the suspension system power bus or across thecentral converter may be used by the entire vehicle as an energy buffermechanism.

Exemplary system conditions that may be determined from the power busvoltage level are shown in FIG. 2B that shows the full extent of thepower bus divided into operating condition ranges. Normal operatingrange conditions may include net regeneration and bias low energy. Whenthe power bus level signals that the system is in a mode of netregeneration, a suspension control system may activate functions such assupplying voltage to the vehicle primary power bus. A bias low energysystem condition may indicate to a suspension control system thatavailable energy reserves are being taxed, so preliminary measures toconserve energy consumption may be activated. In an example ofpreliminary energy consumption measures, wheel event response thresholdsmay be biased toward reducing energy demand. Alternatively, when a biaslow energy system condition is detected, energy may be requested fromthe primary vehicle power bus to supplement the suspension system poweravailability. Above a normal operating range or mode may be a load dumprange. This may be indicative of the suspension system re-generatingexcess energy that is so high that it cannot be passed to the vehicleprimary power bus, so that there is a need for it to be shunted off. Asuspension system controller, such as a corner control for a vehiclewheel, may detect this system condition and respond accordingly toreduce the amount of energy that is regenerated by the controller'ssuspension actuator sub system. One such response may be to dissipateenergy in the windings of an electric motor in the suspension actuator.Operating modes that are below normal operating range may include faulthandling and recovery modes, and an under-voltage shutdown mode. Inembodiments, operation in a fault handling and recovery mode may signalto the individual corner controllers to take actions to substantiallyreduce energy demand. To the extent that each corner controller may beexperiencing different wheel events, stored energy states, and voltageconditions, the actions taken by each corner controller may vary, and inembodiments different corner controllers may operate in differentoperating modes at any given time. An under-voltage shutdown conditionmay be indicative of an unrecoverable condition in the system (e.g. aloss of vehicle power), a fault in one of the independent cornercontrollers, or a more serious problem with the vehicle (e.g. a wheelhas come off) and the like. The shut down mode may cause the cornercontroller to operate solely as a passive or semi-active damper, ratherthan a fully active system.

DC Voltage Bus Level Defines [Vehicle] Energy System Capacity

As noted above, a suspension system power bus DC voltage level maydefine system conditions. It may also define the energy capacity of thesystem. By monitoring the suspension system power bus, each cornercontroller can be self-informed of how much energy is available forresponding to wheel events and maneuvers. Using the power bus tocommunicate suspension system and/or vehicle energy system capacity alsoprovides safety advantages over separated power and communication buses,because a fault on one of the separated buses could result ininconsistent wheel operation that may lead to a safety concern. By usingvoltage levels of the power bus to signify operational conditions andpower capacity, each corner controller can operate without concern thata corner controller is missing important commands that are beingprovided over a separate communication bus to the other cornercontrollers. In addition, it either eliminates the need for a signalingbus (which requires wiring), or reduces the communication bus bandwidthrequirements.

By providing a common power bus to all, or a plurality of, the cornercontrollers, each corner controller can be safely decoupled from othersthat may experience a fault. In an example, if a corner controllerexperiences a fault that causes the power bus voltage level to besubstantially reduced, the other corner controllers may sense thereduced power bus voltage as an indication of a problematic systemcondition and take appropriate measures to avoid safety issues.Likewise, with each corner controller capable of operating independentlyas well as being tolerant of complete power failure, even under severepower supply malfunction, the corner controllers still do the rightthings to ensure reasonable suspension operation.

By having a carefully managed coupling between the suspension power busand the vehicle power bus as shown in FIGS. 2A and 2C, a power converterthat facilitates converting vehicle power to suspension system power maybe simplified because the information about each power bus that isnecessary for managing the coupling can be obtained from the suspensionsystem power bus. Information such as, vehicle battery status (e.g.under charged, over charged, etc), suspension system energy storagestatus, and the like can be determined by monitoring the voltage levelof the suspension system DC power bus.

The fluctuating voltage bus topology provides for an effective energystorage architecture. Super capacitors may be placed on the bus, and thebus voltage would define the amount of available energy in thecapacitors (higher voltage bus has more energy in the capacitors). Byreading the current voltage level, each corner controller knows thecurrent energy storage level in the capacitors and can adapt suspensioncontrol dynamics based on this knowledge. By way of illustration, for aDC bus that is allowed to fluctuate between 38V and 50V, with totalstorage capacitance C, the amount of available energy (neglectinglosses) is:Potential Energy=V*C*(50){circumflex over ( )}2−½*C*(38){circumflex over( )}=528*C

Using this calculation or similar, the corner controllers are able toadapt algorithms to take into account the limited storage capacity,along with the static current capacity of a central DC/DC converter tosupply continuous energy.

Corner Controller Damper which can have a Voltage Failure Tolerant SmartValve

An active vehicle suspension system may include a vehicle suspensioncorner controller at each wheel. As noted above, this configuration maybenefit safety of the vehicle in that each corner controller may monitora suspension system power bus to determine actions to take to maintainsafe operation. In the event of power loss to a corner controller, it isimportant that the suspension system for the controller that is affectedcontinues to operate safely. In addition to power loss, as depicted inFIG. 2H (disconnect failure mode), the suspension of any wheel must betolerant of a short-circuit failure mode condition as depicted in FIG.2I. In either condition, the suspension system (e.g. actuator, hydraulicmotor, and related components) must operate at least as good as apassive suspension component so that the vehicle can be operated safelywith wheel movement mitigated enough to prevent loss of criticalfunctions such as steering, braking, and the like. Therefore a cornercontroller of an active suspension system may default to operating asindependent passive suspension (e.g. traditional damper) for each wheelwhen a power failure is detected. In another embodiment, each cornercontroller of an active suspension system may default to operating asindependent semi-active suspension (e.g. adaptive damper) for each wheelwhen a power failure is detected. Such as system might utilizeself-powered adaptive damper technology disclosed earlier.

Super Capacitor Use in a Vehicle Active Suspension System

An active vehicle suspension system may include one or more highcapacity electrical capacitors (e.g. super capacitors) configured todeliver transient energy to the actuators of the active suspension. Inembodiments this may include running a bank of capacitors at a voltagehigher than the vehicle's electrical system voltage, in order to powerthe actuators. In at least some modes, energy is able to flow eitherdirectly or through a power converter from the vehicle's electricalsystem to charge the super capacitors. In embodiments the supercapacitors may be placed across the input and output of a DC/DCconverter, such that the negative terminal is on the vehicle electricalsystem positive voltage, which is connected to one side of the DC/DCconverter, and the positive terminal is on the other side of the DC/DCconverter. For example, for a unidirectional or bidirectional 12V to 42Vconverter, the capacitors may be placed from the positive 12V to thepositive 42V. By doing this, the voltage across the capacitors is only30V instead of 42V if placed across 42V and ground. This allows fewercapacitors to be used, as super capacitors generally have voltageratings between 1V and 3V, requiring several capacitors to be placed inseries when run at higher voltages. This configuration allows very costeffective use of energy storage in an active suspension system.

Vehicular Bi-Directional DC/DC Converter & System

An active suspension system power sub-system may derive primary energyfrom a vehicle electrical system (e.g. from the vehicle car battery)through a DC-DC voltage converter that provides voltages higher than atraditional car battery (e.g. +12 VDC) to operate the suspension systemcomponents, such as 3-phase electric motors, and the like. As notedabove herein, such a DC-DC voltage converter may be a bi-directionalconverter that allows energy in the vehicle electrical power bus to beconverted to a higher voltage for use by the suspension system and thehigher voltage energy in the suspension system bus to be converted tovehicle battery compatible voltage and provided into the vehicle primaryelectrical system. In addition, energy storage for an active vehiclesuspension system may be beneficial to facilitating on-demand energydelivery, such as to respond to a wheel event, without overly taxing thevehicle electrical system. To add energy storage, one may add a highstorage capacity capacitor to the higher (e.g. +48V) suspension powerbus; however, capacitors (e.g. super caps) that are capable of storing asubstantial amount of energy while providing a nominal +48V are verylarge and expensive. To provide a nominal 48V a capacitor that canhandle as much as 60V may be required, increasing the size and cost evenfurther. Therefore, rather than referencing a storage capacitor (e.g.super cap) for the suspension system +48V power bus to ground, it may bereferenced to the vehicle primary electrical system that is controlledby the vehicle battery, which his typically +12 VDC. This reduces thevoltage range that is required to be supported by the super capacitor byas much as, for example, 12V. The result is a smaller, lower costcapacitor can be used to provide energy storage.

Advantages of this approach may include reducing the number of cells inthe super capacitor, which reduces cost and size, and eases theimpedance requirements of the capacitor, because impedance of a supercap may be proportional to the number of series cells. The result ismore efficient charging and discharging of the super capacitor. Anotheradvantage is that because the super capacitor is referenced to thevehicle battery voltage, at startup of the suspension system powersystem (e.g. the DC/DC converter and central controller of FIG. A) thesuper capacitor is already at the battery voltage, so this minimizesinrush of current to the super capacitor. Minimizing in-rush isdifficult with super capacitors, so having this initial voltage on thesuper capacitor is highly beneficial, especially with regard to theparticular requirements of automotive applications, because it preventshigh current draw from the vehicle battery when the suspension system isfirst energized.

The vehicular bidirectional DC/DC converter and system allows energy toflow in both directions, and the energy flow sizing may be imbalancedfor each direction. For example, in the case of a configurationcomprising directionally opposed buck and boost converters, eachconverter may be sized to handle a different amount of power. For a 12Vto 46V system, for example, the continuous power capability from 12V to46V may be 1 kilowatt, while from 46V to 12V in the reverse direction itmay only be 100 watts. This asymmetrical sizing may save cost,complexity, and space. These factors are especially important inautomotive applications. In embodiments the DC/DC converter is used asan energy buffer/power management system, and the input and outputvoltages are roughly equivalent (e.g. 12V to 12V converter). Inembodiments the DC/DC converter may contain one side that has a voltagethat fluctuates, for example, between 24V and 60V. or 300V and 450V.This may be controlled as part of a larger energy management system asdiscussed earlier.

While one application of the vehicular bidirectional DC/DC converter isfor active suspensions, it may also be used to power other devices suchas in a vehicle electrical system (e.g. electric power steering, ABSbrakes, etc.) or on a hybrid vehicle.

Another aspect of the inventive methods and systems are the ability foran external energy management control signal to regulate power. Inembodiments an external CAN bus signal from the vehicle is able to sendcommands to the DC/DC converter in order to dynamically manage andchange directional power limits in each direction, or to downloadvoltage limits and charge curves.

Control Topology of an Active Suspension Including a Three-Phase ACController on a Local Power Bus (e.g. 48 VDC) which can be Coupled to aVehicle Power Bus

Control Topology of an Active Suspension Including a Processor-BasedController Per Wheel

Control topologies for active suspension systems may vary based ondesign objectives and performance goals. For an active suspension systembased around a damper system coupled to a hydraulic motor and co-axialelectric motor/generator, using 3-phase AC control of the electricmotor/generator may provide high performance over a wide range ofoperating conditions. Three-phase control of a vehicle suspensionactuator may be combined with the power bus features and functionsdescribed above and elsewhere herein of a local suspension system powerbus that is controllably coupled to a vehicle primary electrical powerbus. Such a combination may provide robust suspension control whilegaining the safety and startup advantages described herein. Also,configuring each wheel with its own processor-based controller (e.g. acorner controller) and individual 3-phase AC circuit may allow precisewheel-by-wheel response to road conditions. Certain 3-phase AC circuitconfigurations at each corner may also improve safety and fail-safeoperation in that a 3-phase AC bridge at each corner controller (e.g. asshown in FIG. 2J) may convert operation to a power rectifier when thesuspension system power bus is lost (e.g. as shown in FIG. 2K).

In addition to individual wheel control through a 3-phase AC generator,a control topology that includes coordination of the corner controllersmay be suitable for vehicle dynamics control. FIG. 2F depicts adistributed vehicle dynamics control topology in which vehicle systemsmay coordinate with vehicle suspension via a gateway that propagatesvehicle dynamics information to the corner controllers in a distributedfashion. Each corner controller may sense local wheel conditions andshare this information with other corner controllers who may share thisinformation and locally sensed wheel information with other controllersand through the gateway with other vehicle systems. In this way eachcontroller may adapt its functioning based on locally sensed wheelevents, as well as based on neighboring corner controller wheel events,and even based on vehicle-level events from external ECUs (e.g. theelectronic stability control system).

A second control topology that may be suitable for vehicle dynamicscontrol is depicted in FIG. 2G as a centralized vehicle dynamicsvariant. In a centralized control topology, each wheel may share localwheel event information with a central controller that may also exchangevehicle dynamics-related information with other systems of the vehicle.The central controller may provide vehicle dynamics information, such asdetails about vehicle systems, other wheel events, wheel-specificinstructions, and the like directly to the individual corner controllersto operate and maintain desired vehicle suspension operation.

While in many embodiments a rotary electric motor is controlled, in someembodiments the electronics described may control a linear electricmotor. In other embodiments an electric motor may be controlled, whichin turn, controls a ball screw mechanism or a hydraulic actuationmechanism. In other embodiments the controllers are coupled with valvecontrollers that operate electronically controlled valves.

In embodiments, a normally closed relay and connected damping device(such as a short circuit or resistor) is added to the system such thatwhen power is applied it is held open. When power is lost to the relay,the electric motor of an active suspension is shunted in order toprovide damping force. This is especially advantageous in order toproduce a graceful failure mode. By selecting the appropriate resistancefor the relay device, the damping curve of the actuator in failure modemay be tuned. Other configurations may use a low energy relay or otherlatching device. In some configurations a relay without a damping devicemay be used in order to disconnect the actuator from an electrical load(to create low damping force). Since hydraulic active suspension systemsmay be tuned to have passive damping forces at open circuit on theelectric motor, even in open circuit a damping force may be created andtuned.

Turn now to the figures and initially FIG. 2A, which shows the blockdiagram of an embodiment of an active suspension that includes fourGenShock Smart Valves, four corner controllers, one central controller,the vehicle electrical bus, various feedback sensors, the DC powerdistribution bus, and the communications and control bus. While theembodiment of FIG. 2A may contain four corner controllers, the inventivemethods and systems are not limited in this regard. It is equallyapplicable to vehicles with more or less than two axles and fourcorners.

In this embodiment, the GenShock Smart Valve 2-1 includes a brushless DCmotor (BLDC) and an integrated position sensor 2-6 that providesfeedback on the position of the BLDC motor rotor which allows the cornercontroller 2-2 to control the torque and speed of the BLDC motor. Thehydraulic motor (not shown) is directly coupled to the BLDC motor andprovides hydraulic pressure and flow in direct proportion to the motortorque and speed.

The corner controller 2-2 includes a power control module 2-3 thatbi-directionally converts the DC bus voltage and current to AC phasevoltages and currents provided to the BLDC motor. When properly alignedwith the motor rotor using feedback from the position sensor 2-2, theseAC currents control the torque in the BLDC motor. When the motor torqueis in the same direction (clockwise CW or counterclockwise CCW) as themotor rotation, power flows from the DC bus, through the BLDC motor andinto balance of the GenShock system. When the motor torque is oppositethe direction of motor rotation, power flows out of the GenShock system,through the BLDC motor and into the DC bus. Thus, the corner controllercontrols the BLDC motor (and the coupled hydraulic motor) in all fourquadrants: active CW, active CCW, regen CW and regen CCW.

The corner controller 2-2 also takes feedback from local corner sensors2-5 and by adjusting the torque in the BLDC motor, controls the wheelacceleration, body acceleration, and/or the relative spacing of thewheel within the corner of the vehicle (damper position) or dampervelocity. In this case, use of the word damper includes a fullfour-quadrant capable active actuator. In one embodiment, the controlalgorithms to perform these functions reside in the distributed vehicledynamics module 2-4 within each corner controller.

The DC power bus 2-8 distributes power to and from the cornercontrollers and in turn either draws power from, or provides power to,the vehicle electrical bus 2-16 as needed. To accomplish this, thecentral controller 2-9 may contain a bidirectional DC/DC converter 2-10capable of flowing power between a range of DC voltages. In 12Vvehicles, the voltage on the vehicle electrical bus typically rangesfrom 10 to 16 VDC depending on vehicle loads, alternator power and thestate of charge (SOC) of the main vehicle battery. In some applications,the voltage range could be higher or lower, depending on the nominalvoltage and type of the battery used. The DC power bus is typicallyoperated at voltages higher than the vehicle electrical bus to reducethe current requirements, improve efficiency, allow smaller and lessexpensive wire gauge in the distribution bus, and to allow power sharingbetween the corner controllers. While reference is made to a 12Velectrical system, the present inventive methods and systems are notlimited in this regard (e.g. it can function with 24V systems on trucksand buses, or 42-48V systems in future passenger cars).

The central controller 2-9 also may contain energy storage components2-11 (supercapacitors, high power-density batteries, or the like) tobuffer some of the power requirements of the corners and reduce the peakpower and current drawn from, or returned to, the vehicle electricalbus. This allows the active suspension to have a minimum negative impacton the vehicle electrical system, which eases system integration. Thebidirectional DC/DC converter 2-10 can control the voltage on the DCpower bus over a wide range depending on the requirements of theapplication and to allow energy to be efficiently stored in the energystorage component. Supercapacitors, regular capacitors and somebatteries require voltage compliance to store energy effectively.

Optionally, the central controller can also contain a centralizedvehicle dynamics controller 2-12 to coordinate the operations of thecorner controllers and respond to overall vehicle conditions and eventssuch as steering, throttle and brake inputs, yaw rate, longitudinal andlateral acceleration, etc. Alternatively, these control responsibilitiescan reside in the distributed vehicle dynamics controllers 2-4 embeddedin the corner controllers. As yet another centralized alternative, thecontrol responsibilities can reside in an external ECU on the vehiclesuch as a chassis controller. In both centralized and decentralizedembodiments, vehicle conditions and event information is communicatedfrom the vehicle chassis or other ECU 2-14 to the active suspensioncontrol system via a vehicle communications bus 2-15 and acommunications gateway 2-13. This gateway can be part of the centralcontroller 2-9, the central vehicle dynamics controller 2-12, thevehicle ECU 2-14 or some other place within the vehicle. In someembodiments a hybrid control strategy uses local processing of highfrequency calculations directly in the distributed vehicle dynamicscontrollers 2-4, and low frequency calculations in a centralizedcontroller.

In some embodiments the corner controllers also communicate to eachother and to the central controller, gateway, and optionally, thecentral vehicle dynamics controller, over a private communications bus2-7. This bus can be serial (CAN, FlexRay, RS485. Ethernet, etc.),optical, or parallel. By separating the private communications bus fromthe vehicle communications bus with a gateway, the active suspensionsystem can utilize the full bandwidth of the private communications buswithout affecting the communications bandwidth required for othervehicle functions.

At times when the vehicle electrical bus 2-16 cannot accept any morepower and the energy storage component of the active suspension 2-11 isalready at full capacity, the central control can dump any regeneratedpower into a dump resistor 2-17 or similar dump load. Dumpingregenerated power does not otherwise adversely affect the performance ofthe active suspension system.

In the embodiment shown in FIG. 2A, the energy storage component 2-11 isa capacitor or preferably, a supercapacitor. The DC power bus 2-8 isallowed to float up and down within a voltage control range to store andextract energy from the supercapacitor. Since the energy stored in acapacitor is a function of the square of its operating voltage, thecentral controller and all of the connected corner controllers can knowthe energy storage state of the active suspension control system bymeasuring the DC voltage on the bus. FIG. 2B shows the operating stateof the active suspension system as a function of the voltage on the DCpower bus. The normal operating range 2-23 of the DC power bus 2-8 isbetween Vhigh 2-19 and Vlow 2-21 with a control setpoint, Vnom 2-20,between these two limits. When one or more of the corner controllersrequire power, they take what they need from the DC bus and the busvoltage decreases below Vnom 2-20 because energy is removed. Thebidirectional DC/DC converter 2-10 reacts by taking power from thevehicle electrical bus to try to keep the DC bus voltage near Vnom 2-20.But since the entire active suspension control system can functionnormally down to Vlow 2-21, the DC/DC converter lets the DC power busvoltage sag a bit and some of the energy is instead taken from theenergy storage supercapacitor. This strategy reduces the peak power loadon the vehicle electrical bus.

Conversely, when the corner controllers are regenerating and providingpower to the DC bus 2-8, the bus voltage increases above Vnom 2-20 andthe bidirectional DC/DC converter 2-10 reacts by putting power into thevehicle electrical bus. Again, since the active suspension system canoperate normally up to Vhigh 2-19, the DC/DC converter lets the DC busvoltage peak a bit and some of the energy is instead put into thesupercapacitor 2-11. This strategy reduces the peak power flowing fromthe corners into the vehicle electrical bus. This is desirable incertain circumstances as delivering high currents of regenerated energyto a lead acid battery can cause overvoltage and damaging conditions.

The bidirectional DC/DC converter 2-10 limits the peak current into orout of the vehicle electrical bus to avoid disturbing other systemswithin the vehicle. If the power requirements of the corner controllersare beyond the capability of the DC/DC converter at its current limit,the DC bus voltage may either go below Vlow 2-21 as the supercapacitorenergy is depleted or above Vhigh 2-19 as the supercapacitor energyreaches its maximum.

Below Vlow 2-21, the corner controllers 2-2 react by reducing theirpower requirements, thus allowing the DC/DC converter to recharge thesupercapacitor. Above Vhigh 2-19, the central controller reacts bydumping the excess energy into its load dump 2-17 and the cornercontrollers 2-2 react by changing the operation of the motor toregenerate less power.

If the total regenerated power is so high that the DC bus voltagecontinues to rise above Vhigh 2-19 and it reaches the Overvoltage (0V)threshold 2-18, the corner controllers may react by either lowering themotor torque, in the limit all the way to zero, disabling the powercontrol module 2-3, shorting the windings of the BLDC motor, or somecombination of these actions. Conversely, if the power demand is sogreat that the DC bus voltage continues to drop below Vlow 2-21, thesystem may continue to function at reduced power until the system eitherrecovers or the Undervoltage (UV) threshold 2-22 is crossed at whichpoint the corner controllers may shut down.

The setpoint Vnom 2-20 can be dynamically adjusted by the vehicledynamics controller to allow more headroom for regenerative events ormore capacity for active (power consuming) events. Additionally, thevehicle dynamics controller can always know how much energy is left forcertain maneuvers by measuring the DC bus voltage, or how much headroomis available for storing regenerated energy.

Throughout the entire voltage range show in FIG. 2B, the cornercontrollers 2-2 and the bidirectional DC/DC converter 2-10 are able tofunction normally and react correctly to out of range conditions withoutthe need for communications other than via the DC bus voltage level.This feature adds robustness and safety to the active suspension controlsystem.

Energy storage is a major cost driver for an active suspension controlsystem. Series strings of supercapacitors are an appealing option butonly for intermediate voltages (less than 100V) where the number ofcells is reasonable and the cell voltage can be de-rated properly foroperating temperature. FIG. 2C shows six embodiments of supercapacitorstrings 2-11 integrated with a bidirectional DC/DC converter 2-10.

The first embodiment 2-24 has the supercapacitor string on the DC buswhere the voltage compliance is large but the voltage across the stringis also high. It requires 20 cells in series at 2.5V/cell.

The second embodiment 2-25 has the supercapacitor string on the vehicleelectrical bus in parallel with the vehicle battery where the voltagecompliance is defined by the vehicle alternator, battery and loads, andis therefore low, but the voltage across the string is also low. Itrequires only 6 to 7 cells but the cells must be much larger capacitancethan the first embodiment 2-24.

The third embodiment 2-26 has the supercapacitor string in series withthe vehicle electrical bus 2-16. This topology can have large voltagecompliance but generally works in applications where the current in thesupercapacitor string averages to zero. Otherwise uncorrected, thesupercapacitor string voltage may drift toward zero or overvoltage.

The fourth embodiment 2-27 has the supercapacitor string in series withthe output of the DC/DC converter. Like embodiment 2-26, this topologyrequires that the current in the supercapacitor string averages to zero,which limits its application.

The fifth embodiment 2-28 has the supercapacitor string across the DC/DCconverter between the vehicle electrical bus 2-16 and the DC power bus2-8. This topology is functionally similar to 2-24, but it reduces thenumber of cells required from 20 to 16 by referencing the supercapacitorstring to the vehicle electrical bus rather than chassis ground,reducing the string voltage requirement by at least 10 V (the minimumbattery voltage.)

The sixth embodiment 2-29 solves the average DC bus current limitationof embodiment 2-27 by adding an auxiliary DC/DC converter 2-30 to assurethat the supercapacitor string current averages to zero even when the DCbus current does not average to zero.

Other logical combinations of these embodiments, such as adding theauxiliary DC/DC converter 2-30 to embodiment 2-26 are also possible. Thebest topology for a specific application primarily depends on the costof supercapacitors as compared to power electronics and on theinstallation space available. Additionally, alternative energy storagedevices than supercapacitors such as lithium ion batteries may be usedin the same or similar configurations as those disclosed here.

FIG. 2D shows an active suspension system for a 4-wheeled vehicle 2-31with a corner controller 2-2 at each wheel 2-33, each activelycontrolling a GenShock damper 2-32 at that wheel. A communications bus2-7 between the corners enables coordination between the controllers.While the embodiment of FIG. 2D is for a 4-wheeled vehicle the inventivemethods and systems are not limited in this regard. It is equallyapplicable to vehicles with more or less than four wheels (includingmotorcycles, light trucks, vans, commercial trucks, cargo trailers,trains, boats, multi-wheeled and tracked military vehicles, and othermoving vehicles).

FIG. 2E shows a corner controller embodiment 2-2 comprising amicrocontroller unit (MCU) 2-36 with a built-in communications bus 2-7port, an analog to digital converter (ADC) 2-35 for reading sensorsvalues 2-5 into the MCU and a power control module 2-34 producing one ormore actuator control signals 2-37. The ADC may be a functional block ofthe MCU or an external component.

FIG. 2F shows an active suspension system with distributed vehicledynamics controllers 2-36 in each corner controller 2-2. The distributedcontrollers are not only responsible for control of their local wheeland corner of the vehicle but also for coordination between corners andfor responding to overall vehicle conditions and events such assteering, throttle and brake inputs, yaw rate, longitudinal and lateralacceleration, etc. The corner controllers communicate with each otherover a private communications bus 2-7 and with the vehicle through agateway 2-13 that bridges to the vehicle communications bus 2-15.

FIG. 2G shows an active suspension system with a centralized vehicledynamics controller 2-38 that is responsible for coordination betweencorners and for responding to events affecting the entire vehicle, suchas driver inputs (steering, throttle and brake inputs), or vehiclemotions caused by other external influences (yaw rate, longitudinal andlateral acceleration, etc). The central vehicle dynamics controller 2-38communicates with each corner controller over a private communicationsbus 2-7 and with the vehicle over the vehicle communications bus 2-15.In this embodiment, the central vehicle dynamics controller alsofunctions as the gateway to separate the private communications bus 2-7from the vehicle communications bus 2-15.

FIG. 2H depicts the “DC bus disconnected” failure mode for a corner. Thecorner controller 2-2 is located in close proximity to a valve 2-40mounted on a damper 2-39. The valve may contain a permanent magnet motorcoupled to a hydraulic pump. The corner controller powers the motorphase windings 2-41 and is connected to the DC power bus 2-8 and tochassis ground 2-42. The failure mode is when a break 2-43 in the DCpower bus wiring causes the DC bus to become an open circuit. In thisfailure mode, the corner controller can still provide controlled dampingby using regenerated energy to keep the DC bus capacitor 2-45 (see FIG.2J) in its normal operation range 2-23. The corner controller knows thatthe DC bus is disconnected since as the impedance has increaseddramatically.

FIG. 2I depicts the “DC bus shorted” failure mode for a corner. Thecorner controller 2-2 is located in close proximity to a valve 2-40mounted on a damper 2-39. The valve may contain a permanent magnet motorcoupled to a hydraulic pump. The corner controller powers the motorphase windings 2-41 and is connected to the DC power bus 2-8 and tochassis ground 2-42. The failure mode is when a short occurs between theDC bus 2-8 and chassis ground 2-44 forcing the DC bus to zero volts andimmediately causing an undervoltage 2-22 shutdown of the cornercontroller. In this failure mode, the flyback diodes 2-49 & 2-50 (seeFIG. 2J) in the 3-phase bridge of the power control module 2-3inherently act as a 3-phase rectifier, effectively shorting the windingsof the motor and providing short circuit damping.

FIG. 2J shows an embodiment of the power control module 2-3 in a cornercontroller 2-2. It comprises of six MOSFETs 2-47 & 2-48 and six flybackdiodes 2-49 & 2-50 connected in a 3-phase bridge configuration with apositive DC rail 2-46 and negative DC rail 2-42. The power controlmodule connects to a 3-phase motor 2-55 via three phase leads: 2-51,2-52 & 2-53. At minimum two of the phase leads have a current sensor2-54 to enable feedback control of the motor winding currents and hencethe motor torque. The third current sensor shown is optional and adds aredundant current measurement for robustness and for short circuit faultdetection. A DC bus cap is connected across the positive and negative DCrails to provide ripple current filtering and holdup time for controlcircuitry powered from the DC bus.

FIG. 2K shows the same power control module as in FIG. 2J during the “DCbus shorted” failure mode for a corner shown in FIG. 2I. Due toundervoltage shutdown and loss of control power, the MOSFETs 2-47 & 2-48are off and the current sensors 2-54 are inoperable, so neither isshown. In this failure mode, the flyback diodes 2-49 & 2-50 in the3-phase bridge inherently act as a 3-phase rectifier, effectivelyshorting the windings of the motor and providing short circuit damping.

Position Sensor

In certain types of regenerative, active/semi-active dampers, anelectric motor is used to provide torque and speed to a hydraulic motorto provide force and velocity to a hydraulic damper, and conversely, thehydraulic motor is used to back-drive the electric motor as a generatorto produce electricity from the force and velocity inputted into thedamper (as shown in FIG. 3-A). For reasons of performance anddurability, these electric motors are of the BLDC type and are mountedinside the damper body, encased in the working fluid under pressure. Inorder to provide adequate damper performance, accurate control of thetorque and speed of the BLDC motor is required which may require arotary position sensor for commutation. Although rotary position sensorsfor BLDC motor commutation/control currently exist, the application foruse in a regenerative, active/semi-active damper valve is particularlychallenging as the BLDC motor is mounted inside of the damper body whereit is encased in the working fluid under high pressures. A device toimprove the control and system feedback of a hydraulic active dampervalve by sensing the rotational position of the BLDC motor is disclosed

Current rotary position sensors are sensitive devices that often cannotbe subjected to hydraulic fluid under pressure. It is thereforenecessary to shield the rotary sensor from the hydraulic fluid pressurewhile not impeding its ability to accurately sense position.

Electric Generator Rotor Position Sensing

Methods and systems described herein may relate to sensing an electricgenerator rotor position. An electric generator may be applied in anactive suspension system to work cooperatively with a hydraulic motor tocontrol movement of a damper in a vehicle wheel suspension actuator. Asdescribed elsewhere herein, the electric generator may be co-axiallydisposed with the hydraulic motor and may generate electricity inresponse to the rotation of the hydraulic motor, while also facilitatingrotational control of the hydraulic motor by applying torque. To deliverrobust suspension performance over a wide range of wheel events, it maybe desirable to precisely control the electric generator. To achieveprecise control, precise rotor position information may be needed.

In particular, determining the position of the rotor relative to thestator (the windings) is important to precisely control currents passingthrough the windings based on the rotor position. To precisely anddynamically control the currents through the windings depending on wherethe rotor is in its rotation, what direction it is turning, itsvelocity, and acceleration, a fairly precise reading of rotor positionis required. To achieve precisely determining the rotor position, asensor is used. By applying position determination algorithms that aredescribed below, a low cost sensor (e.g. with accuracy of 1 degree) maybe used.

In some configurations of an active suspension system described herein,portions of electric generator may be submerged in hydraulic fluid. Thismay present challenges to sensing a precise position of the rotor.Therefore, a magnetic target attached on the rotor shaft may be detectedby a sensor disposed so that it is isolated from the hydraulic fluid.One such arrangement may include disposing a sensor on a dry side of adiaphragm that separates the fluid from the sensor. Because magneticflux passes through various materials, such as a nylon diaphragm, it ispossible to read the rotor position while keeping the sensor out of thefluid.

While a low cost magnetic sensor may provide one-degree resolution withone to two degrees of linearity, which may be sufficient simply fordetermining rotor position, to precisely control the currents flowingthrough the windings, additional information about the rotor may beneeded, such as acceleration of the rotor. One approach would be to usea more accurate sensor, although this increases costs and may not evenbe practical given the rotor is immersed in fluid.

Therefore, a filter that correlates velocity with position wasdeveloped. The filter may perform notch filtering with interpolation ofany filtered positions. By performing notch filtering, harmonics of thefiltered frequency are also filtered out, thereby improving results. Byusing a combination of filtering, pattern sensing, and on-line autocalibration, precise calibration steps during production or deploymentare eliminated, thereby reducing cost, complexity, and service issues.

Methods and systems of rotor position sensing in an active suspensionsystem may include magnetically sensing electric generator rotorposition of a fluid immersed electric generator shaft through adiaphragm. Other methods and systems may include processing the sensedposition data to determine rotor acceleration with a low-cost magneticsensor.

Magnetically Sensing Electric Generator Rotor Position Through aDiaphragm

In hydraulic systems where a motor such as a BLDC motor is disposed influid, it may be desirable to sense rotor position in order to controlthe windings of the motor for commutation (e.g. at low velocities wheresensorless control is prone to errors), or to provide needed feedbacksensory information. In embodiments, the rotor of the motor is disposedin a fluid, with the shaft comprising a polarized magnet co-axial withthe shaft of the rotor and at the top of it. A non-magnetic diaphragmseparates the fluid medium from the non-fluid environment at the top ofthe shaft. A Hall effect sensor and/or magnetic rotary encoder arerigidly mounted on the dry side of the diaphragm, oriented so as tomeasure the magnetic flux from the rotor-mounted magnet.

Sensing Rotor Position of a Fluid Immersed Electric Motor Shaft in anActive Suspension

The system similar to that described in the previous paragraph, whichcan sense rotary position or rotary velocity of a shaft in a fluidmedium, without the need for electronics to be placed within fluid orwithout the need for wire pass-throughs, may be applied to an activesuspension system. Active suspensions that utilize a hydraulic pump andelectric motor may oftentimes place the electric motor outside theworking fluid of the hydraulic system. However, placing the electricmotor in the fluid may provide significant benefits such as improvedthermal cooling, better packaging, and reduced frictional losses from ashaft seal on the hydraulic motor. For active suspensions, it is oftennecessary to measure rotor position, however. A system as that describedabove allows an active suspension to incorporate rotor measurementwithin a fluid medium in a durable and cost-effective manner.

Rotary Hall Effect Position Sensor

In FIG. 3A a regenerative active/semi-active damper 3-100 that comprisesa side mounted hydraulic regenerative, active/semi-active damper valve3-200, and a monotube damper assembly 3-300, is shown.

In the embodiment of FIGS. 3B and 3C, a rotary Hall effect positionsensor 3-9, that measures the rotational position of a source magnet 3-6and is protected from the working hydraulic fluid 3-14 under pressure,is shown. The Hall effect sensor is adequately shielded from otherexternal magnetic fluxes so as not impair its ability to accuratelysense the position of the magnetic flux of the source magnet.

The Hall effect position sensor is mounted on a PCB 3-11 to which thesensor wires 3-12 are connected. The PCB is supported in a bore in asensor body 3-7. In FIG. 3B, the sensor body and sensor are held inrigid connection to the valve body 3-1. The sensor body is constructedof a magnetic material (such as steel for example) so as to shield thesensor from external unwanted magnetic fluxes (from the BLDC motormagnets 3-3 for example) that may degrade the sensor accuracy. In theembodiment shown, the PCB and the sensor are located coaxially with therotational axis of the BLDC motor 3-2. A sensor shield 3-8 is locatedwithin the sensor body in front of the sensor and is exposed to thehydraulic fluid under pressure in the damper valve. In FIG. 3C thesensor shield is sealed to the sensor body (by means of a hydraulicseal, or mechanical seal, or adhesive etc. 3-10) such that the hydraulicfluid cannot enter the sensor body. The sensor shield is constructed ofa non-magnetic material such as aluminum or an engineered performanceplastic etc., so that the magnetic fluxes of the source magnet can passthrough the sensor shield unimpeded. A small air gap exists between thesensor shield and the sensor so that any deflection of the sensorshield, due to the hydraulic fluid pressure acting on it, does not placeany load onto the sensor itself. A source magnet is located in a magnetholder 3-5 that locates the source magnet coaxially with the BLDC motorrotational axis and the sensor axis, and in close axial proximity to thesensor shield. The source magnet and magnet holder are operativelyconnected to the BLDC motor rotor. The sensor shield is constructed sothat it has a thin wall section that allows the face of the sourcemagnet to be located close to the working face of the sensor so as toprovide sufficient magnetic flux strength to penetrate the sensor so asto provide accurate position signal. The use of a high performanceplastic for the sensor shield (such as PEEK for example), allows for asufficiently thin wall section to be used while having the structuralintegrity to withstand the hydraulic pressure placed upon it and withoutdegrading the source magnetic flux strength. The source magnet holder isconstructed of a non-magnetic material, such as aluminum or anengineered performance plastic etc. so as not to degrade the sourcemagnetic flux strength. The sensor body is sealed to the valve housing(by means of a hydraulic seal, mechanical seal, or adhesive etc. 3-13)so as to allow the sensor wires to pass through the valve housingwithout allowing any leakage of the hydraulic fluid. The sensor wiresare sealed to the sensor body (by means of a hydraulic seal, mechanicalseal, or adhesive etc. 3-16) so as to protect the Hall effect positionsensor from the environment.

Although the embodiment shown in FIG. 3A discloses the sensor bodyassembly and source magnet coaxial with the rotational axis of the BLDCmotor, the sensor body assembly could be eccentric to this axis and thesource magnet could be of the annular type and offer the samefunctionality as the embodiment shown in FIG. 3A.

In the alternative embodiment of FIGS. 3D and 3E a regenerative adaptive(either active or semi-active) damper 3-400 that comprises an in-linemounted hydraulic regenerative, active/semi-active damper valve, in amonotube damper assembly is shown. The operation of the rotary Halleffect position sensor 3-9 is as described in the embodiment of FIGS. 3Band 3C, except as described below.

Referring to FIG. 3E, in this embodiment the floating piston 3-18 andaccumulator chamber 3-19 are housed in the damper body 3-17 directlybehind the electric generator/motor 3-2. The accumulator chamber 3-19may contain a gas under pressure. The sensor body 3-7 is rigidlyconnected to the damper body 3-17 and may contain a journal diameterthat passes through the floating piston 3-18 and into the accumulatorchamber 3-19. The floating piston slides on this journal and may containa seal (not shown) to prevent leakage across the floating piston fromthe pressurized gas in the accumulator chamber. A seal 3-13 prevents gasleaking past the connection between sensor body 3-7 and the damper body3-17. The sensor wires 3-12 pass through a central bore in the sensorbody 3-7 and out of the damper body. A seal 3-16 prevents the ingress ofcontaminants into the sensor cavity 3-15.

Optical Rotary Position Sensor

In an arrangement similar to the embodiment of the Hall effect rotaryposition sensor, an alternative embodiment is to use an optical rotaryposition sensor that measures the rotational position of a reflectivedisc which is protected from the working hydraulic fluid under pressurein a similar manner to that described in the embodiment of FIG. 3C,wherein the optical rotary position sensor comprises of a lighttransmitter/receiver and a reflective disc.

The light transmitter/receiver is mounted on a PCB to which the sensorwires are connected. The PCB is supported in a bore in a sensor body.The sensor body and light transmitter and receiver are held in rigidconnection to the valve body. In the embodiment disclosed, the PCB andsensor body are located off-axis with the rotational axis of the BLDCmotor. A sensor shield is located within the sensor body in front of thelight transmitter and receiver and is exposed to the hydraulic fluidunder pressure in the damper valve. The sensor shield is sealed to thesensor body (by means of a hydraulic seal, or mechanical seal, oradhesive etc.) such that the hydraulic fluid cannot enter the sensorbody. The sensor shield is constructed of an optically clear materialsuch as an engineered plastic or glass etc, so that the light source canpass through the sensor shield unimpeded. A small air gap exists betweenthe sensor shield and the light transmitter and receiver so that anydeflection of the sensor shield, due to the hydraulic fluid pressureacting on it, does not place any load onto the light transmitter andreceiver itself. A reflective disc is drivingly connected to, andcoaxially with, the BLDC motor, and is located near the lighttransmitter and receiver so that light emitted from the lighttransmitter is reflected back to the light receiver via the opticallyclear sensor shield.

The reflective disc may contain markings so as to produce a reflectedlight signal as the disc rotates, the light transmitter receiver thenreads this signal to determine the BLDC motor position. From thisposition motor speed and acceleration can also be determined. Thewavelength of light source used is such it can pass thru the sensorshield, the oil within the valve and any contaminants contained withinthe oil, unimpeded, so that the light receiver can adequately read thelight signal reflected from the reflective disc.

Although the embodiments of figures one and two refer to an electricmotor rotary position sensor for use in certain types of regenerative,active/semi-active dampers, these embodiments can also be incorporatedinto any electric motor-hydraulic pump/motor arrangement whereby theelectric motor is encased in the working fluid (as in compacthydro-electric power packs etc.), and the inventive methods and systemsare not limited in this regard.

Although the embodiments show the use of a rotary Hall effect positionsensor and optical rotary position sensor, various other types of rotaryposition sensor, such as encoders, potentiometers, fiber optic andresolvers etc. may be accommodated in a similar manner, and the patentis not limited in this regard

Hydraulic Buffer

A device to reduce damper harshness at high speed, low amplitude inputsby mitigating the effects of inertia for passive and semi active dampersas well as hydraulic regenerative, active/semi active dampers.

Conventional passive dampers and semi-active dampers use a combinationof valving to provide the desired force velocity curves for any givenapplication. Although the valve design and spring rates etc. are chosento give the required pressure vs. flow characteristics during steadystate operation, when these valve are placed under highly dynamicoperation, the pressure vs. flow characteristics can change dramaticallydue to inertia effects of the valve(s). Therefore, a damper that hasbeen tuned to provide good damping under low speed events (such as bodyroll and heave, speed bumps etc.) can have harshness at high speed, lowamplitude inputs (such as small road imperfections, raised man holecovers etc.). As the particular valve complexity increases. (such as insemi-active proportional valves or hydraulic regenerative, active/semiactive damper valves) so can their inertia increase, resulting in evenmore undesirable harshness during these high speed, low amplitudeinputs.

The embodiments disclosed a hydraulic buffer is shown to provide aneconomical, method to passively dampen out the high speed, low amplitudeinputs, working in series with the existing damper valving (whether itbe purely passive, semi-active or hydraulic regenerative, active/semiactive).

High Frequency Accumulator for an Active Suspension to Mitigate Effectof Low Energy High Frequency Events on an Active Suspension which canhave an In-Tube Accumulator

An active vehicle suspension system may be configured with a hydraulicactuator coupled to a hydraulic motor to manage wheel events byadjusting the pressure within a damper portion of the actuator. Wheelevents result in some change in position of the damper, which reflectsan increase in pressure/flow that causes the hydraulic motor to spin up.However, some wheel events are high frequency and therefore may exceedthe suspension system's ability to react. Also, high-frequency eventsare often low amplitude and therefore do not require significant energyto mitigate the impact on the vehicle. Therefore, the actuator may beconfigured with an accumulator that can absorb low energy high frequencyevents without resulting in fluid movement toward the hydraulic motor.Such an accumulator may effectively mitigate small energy events withouttriggering hydraulic motor spin-up, thereby providing a measure ofbuffering for small energy high frequency events that the hydraulicmotor may not be able to address. Therefore, methods and systems forhandling wheel events in an active suspension system may includeutilizing a high frequency in-tube accumulator to mitigate the effect oflow energy, high frequency events.

Passive/Semi-Active Configuration

In the embodiment of FIG. 4A a passive monotube damper 4-100 thatcomprises a hydraulic buffer assembly 4-200 in conjunction withconventional passive valving 4-300 is shown.

Referring to FIG. 4B, the hydraulic buffer 4-200 is shown incorporatedinto the piston head 4-3 of the damper 4-100. The hydraulic buffercomprises of bore 4-6 which may contain a floating piston and sealassembly 4-5. Side one of the floating piston and seal assembly 4-5 istoward an oil filled chamber 4-4 that is in fluid communication with thevariable pressure side of the damper 4-18 via an orifice 4-7. In theembodiment shown, the variable pressure side of the damper is therebound chamber, and this pressure varies with the damper force. Theconstant pressure side of the damper is the compression chamber 4-17 andremains constant with respect to damping force because it is in fluidcommunication with an accumulator, via the floating piston assembly 4-19and the pressure may only vary with damper position. It is however,possible for the variable pressure side to be the compression chamberand the constant pressure side to be the rebound chamber, and theinventive methods and systems are not limited in this regard.

Side two of the floating piston and seal assembly 4-5 is toward anaccumulator volume 4-8. The accumulator volume 4-8 is sealed off bymeans of a seal cap 4-10. The accumulator volume provides a spring forcethat forces the floating piston and seal assembly 4-5 toward the oilfilled chamber 4-4. This spring force may be provided by a gas volumeunder pressure, and/or a mechanical spring, such as a compression spring4-9. If a gas chamber under pressure is employed, then the accumulatorvolume may incorporate a gas fill port 4-11. A second spring 4-37 may beused to counter act the spring force from the accumulator 4-8 so thatthe piston is held in a predetermined position, under a pre-load, whenthe damper is in a steady state.

The piston head 4-3 also may contain passive valving 4-13 and 4-14 andflow passages 4-15 and 4-16 which provide the conventional passiverebound and compression damping coefficients. This type of valvearrangement is well known in the prior art.

When damper piston rod 4-2 is subjected to normal rebound strokes thefluid is forced from the rebound chamber 4-28 through the passage 4-16and valving 4-14 to the compression chamber 4-17. The valving 4-14giving a pressure drop that is proportional to the flow rate inaccordance with the desired damping coefficient. A pressure rise in therebound chamber may want to cause fluid flow from the rebound chamber tothe hydraulic buffer oil volume 4-4 however, due to the restriction ofthe orifice 4-7 and the preload applied to the floating piston 4-5 byspring 4-37 and spring 4-9 and or gas precharge in volume 4-8 no fluidflow may occur.

If a high speed, high amplitude rebound stroke is applied to the pistonrod 4-2 the higher pressure that is achieved in the rebound chamber maycause fluid flow from the rebound chamber 4-18 into the hydraulic bufferoil volume 4-4 via the orifice 4-7 as well as through the passage 4-16and valving 4-14 to the compression chamber 4-17. Due to the small gasvolume in the chamber 4-8 and/or the high spring rates of springs 4-37and 4-9 only a relatively small amount of oil may flow into thehydraulic buffer oil volume 4-4 until the pressure in the hydraulicbuffer oil volume 4-4 and the pressure drop across the orifice 4-7equalizes the pressure in the rebound chamber causing the flow into thehydraulic buffer oil volume 4-4 to cease. When the speed of the pistonrod reduces and the pressure in the rebound chamber falls, pressure inthe hydraulic buffer oil volume 4-4 may now be higher than that of therebound chamber and oil may start to flow back out of the hydraulicbuffer oil volume 4-4 into the rebound chamber. Due to the throttlingeffect of the orifice 4-7 the volume of oil may flow out of thehydraulic buffer oil volume. 4-4 at a relatively slow and controlledmanner. Due to the small amount of oil that flows into the hydraulicbuffer oil volume 4-4 compared to the volume of oil that flows throughthe passive valving 4-14 into the compression chamber 4-17 the effect ofthe hydraulic buffer may be negligible under a high speed, highamplitude rebound event.

In a high speed, low amplitude rebound stroke, the velocity of thedamper remains relatively low but may have a high acceleration. Also theamount of displaced volume may be relatively small. The low rod velocitymay generate low fluid flow velocity from the rebound chamber 4-18through the passage 4-16 and valving 4-14 into the compression chamber4-17 whereby the valving should provide low rebound pressures. However,due to the fact that the oil flow is accelerating at a high rate,inertia of the valving 4-14 may cause a pressure spike higher thandesired. This pressure spike may be reflected back into the reboundchamber 4-18 and may cause fluid flow from the rebound chamber 4-18 tothe hydraulic buffer oil volume 4-4 via the orifice 4-7 overcoming thepreload on the floating piston 4-5. Even though there is a high springrate on the floating piston, due to the small gas volume in the chamber4-8 and/or the high spring rates of springs 4-37 and 4-9 a small amountof fluid can be accommodated without much of a pressure rise. Due to thesmall total volume of oil displaced by this event, the small amount ofoil absorbed into the hydraulic buffer oil volume 4-4, may have a bigimpact in reducing the inertia of the oil, and hence its pressure rise.The hydraulic buffer assembly 4-200 may therefore damp out the harshnessnormally felt from these high speed, low amplitude events.

Again, when the speed of the piston rod reduces during this event andthe pressure in the rebound chamber falls below that of the pressure inthe hydraulic buffer oil volume 4-4 fluid may now start to flow back outof the hydraulic buffer oil volume 4-4 into the rebound chamber.

In compression events, the hydraulic buffer assembly 4,200 may operatein the same manner as above, except instead of pressure rises in therebound chamber, there may be pressure drops in the rebound chamber, andinstead of fluid flowing into the hydraulic buffer assembly 4,200 fromthe rebound chamber 4-18 fluid may flow out of the hydraulic bufferassembly 4-200, and into the rebound chamber 4-18.

Although the hydraulic buffer assembly 4,200 is shown as a part of thepiston head of a monotube damper in this embodiment, the hydraulicbuffer assembly can in fact be mounted anywhere on the damper (be it ona monotube, twin tube or triple tube damper, mounted internal, orexternal, to the damper on a passive or semi-active damper) as long asthe orifice 4-7 is in fluid communication with the variable pressurechamber of the damper (as mentioned above), and therefore the patent isnot limited in this regard.

Passive/Semi-Active Configuration.

In the embodiment of FIG. 4C a regenerative active/semi active damperwith a monotube architecture 4-400 that comprises a hydraulic bufferassembly 4-600 in conjunction with a hydraulic regenerative,active/semi-active damper valve 4-500 is shown.

Referring to FIG. 4D, the hydraulic buffer, 4-600, is shown incorporatedinto the piston head 4-23 of the damper 4-400. The hydraulic buffercomprises of bore 4-26 which may contain a floating piston and sealassembly 4-25. Side one of the floating piston and seal assembly 4-25 istoward an oil filled chamber 4-24 that is in fluid communication withthe variable pressure side of the damper 4-34 via an orifice 4-27. Inthe embodiment shown, the variable pressure side of the damper is therebound chamber, and this pressure varies with the damper force. Theconstant pressure side of the damper is the compression chamber 4-33 andremains constant with respect to damping force because it is in fluidcommunication with an accumulator, via the floating piston assembly 4-35and the pressure may only vary with damper position. It is however,possible for the variable pressure side to be the compression chamberand the constant pressure side to be the rebound chamber, and the patentis not limited in this regard.

Side two of the floating piston and seal assembly 4-25 is toward anaccumulator volume 4-28. The accumulator volume 4-28 is sealed off bymeans of a seal cap 4-30. The accumulator volume provides a spring forcethat forces the floating piston and seal assembly 4-25 toward the oilfilled chamber 4-24. This spring force may be provided by a gas volumeunder pressure, and/or a mechanical spring, such as a compression spring4-29. If a gas chamber under pressure is employed, then the accumulatorvolume may incorporate a gas fill port 4-31. A second spring 4-38 may beused to counter act the spring force from the accumulator 4-28 so thatthe piston is held in a predetermined position, under a pre-load, whenthe damper is in a steady state.

The piston head 4-23 also may contain a seal 4-32 that runs in the boreof the pressure tube to resist fluid flow across the piston head fromthe rebound chamber 4-34 to the compression chamber 4-33 and vice versa.

When damper piston rod 4-22 is subjected to normal rebound strokes thefluid is forced from the rebound chamber 4-34 through the a hydraulicregenerative, active/semi active damper valve, 4-500, to the compressionchamber 4-33. The hydraulic regenerative, active/semi active dampervalve 4-500 is controlled so as to increase the pressure in the reboundchamber 4-34 so as to achieve the desired damping coefficient. Apressure rise in the rebound chamber may want to cause fluid flow fromthe rebound chamber to the hydraulic buffer oil volume 4-24, however,due to the restriction of the orifice 4-27 and the preload applied tothe floating piston 4-25 by spring 4-38 and spring 4-29 and/or gasprecharge in volume 4-28 no fluid flow may occur.

If a high speed, high amplitude rebound stroke is applied to the pistonrod 4-22 the higher pressure that is achieved in the rebound chamber maycause fluid flow from the rebound chamber 4-38 into the hydraulic bufferoil volume 4-24 via the orifice 4-27 as well as through the hydraulicregenerative, active/semi active damper valve 4-500 to the compressionchamber 4-33. Due to the small gas volume in the chamber 4-28, and/orthe high spring rates of springs 4-38 and 4-29 only a relatively smallamount of oil may flow into the hydraulic buffer oil volume 4-24 untilthe pressure in the hydraulic buffer oil volume 4-24, and the pressuredrop across the orifice 4-27 equalizes the pressure in the reboundchamber causing the flow into the hydraulic buffer oil volume 4-24 tocease. When the speed of the piston rod reduces and the pressure in therebound chamber falls, pressure in the hydraulic buffer oil volume 4-24may now be higher than that of the rebound chamber and oil may start toflow back out of the hydraulic buffer oil volume 4-24 into the reboundchamber. Due to the throttling effect of the orifice 4-27 the volume ofoil may flow out of the hydraulic buffer oil volume 4-24 at a relativelyslow and controlled manner. Due to the small amount of oil that flowsinto the hydraulic buffer oil volume 4-24 compared to the volume of oilthat flows through the hydraulic regenerative, active/semi active dampervalve 4-500 into the compression chamber 4-33 the effect of thehydraulic buffer may be negligible under a high speed, high amplituderebound event.

In a high speed, low amplitude rebound stroke, the velocity of thedamper remains relatively low but may have a high acceleration. Also theamount of displaced volume may be relatively small. The low rod velocitymay generate low fluid flow velocity from the rebound chamber 4-34through the hydraulic regenerative, active/semi active damper valve4-500 into the compression chamber 4-33, whereby the control on thehydraulic regenerative, active/semi active damper valve 4-500 couldrequire low rebound pressures. However, due to the fact that the oilflow is accelerating at a high rate, inertia hydraulic regenerative,active/semi active damper valve 4-500 may cause a pressure spike higherthan desired. This pressure spike may be reflected back into the reboundchamber 4-34 and may cause fluid flow from the rebound chamber 4-34 tothe hydraulic buffer oil volume 4-24 via the orifice 4-27 overcoming thepreload on the floating piston 4-25. Even though there is a high springrate on the floating piston, due to the small gas volume in the chamber4-28 and/or the high spring rates of springs 4-38 and 4-29 a smallamount of fluid can be accommodated without much of a pressure rise. Dueto the small total volume of oil displaced by this event, the smallamount of oil absorbed into the hydraulic buffer oil volume 4-24 mayhave a big impact in reducing the inertia of the oil, and hence itspressure rise. The hydraulic buffer assembly 4-600 may therefore dampout the harshness normally felt from these high speed, low amplitudeevents.

Again, when the speed of the piston rod reduces during this event andthe pressure in the rebound chamber falls below that of the pressure inthe hydraulic buffer oil volume 4-24 fluid may now start to flow backout of the hydraulic buffer oil volume 4-24 into the rebound chamber.

In compression events, the hydraulic buffer assembly 4-200 may operatein the same manner as above, except instead of pressure rises in therebound chamber, there may be pressure drops in the rebound chamber, andinstead of fluid flowing into the hydraulic buffer assembly 4-600 fromthe rebound chamber 4-34 fluid may flow out of the hydraulic bufferassembly 4-600 and into the rebound chamber 4-34.

Although the hydraulic buffer assembly 4-600 is shown in the piston headof a regenerative active/semi active damper with a monotube architecture4-400 in this embodiment, the hydraulic buffer assembly can in fact bemounted anywhere on the damper (be it on a regenerative active/semiactive damper with monotube, twin tube or triple tube architecture,mounted internal, or external to the damper) as long as the orifice 4-27is in fluid communication with the variable pressure chamber of thedamper (as mentioned above), therefore the inventive methods and systemsare not limited in this regard.

Passive Valving

An active suspension system, such as the system described herein thatincorporates electric motor control of a hydraulic pump, may benefitfrom passive valving that may act as a safety feature. While an activesuspension system may be configured to handle a wide range of wheelevents, pressure buildup of hydraulic fluid may exceed a thresholdbeyond which components of the suspension system may fail or becomedamaged. Therefore, passive valving, such as a rebound throttle valve, acompression throttle valve, a blow-off valve, and the like may beconfigured into the hydraulic fluid flow tubes of the suspension system.

Methods and systems of passive valving in a hydraulic active suspensionsystem may include configuring the system with separate rebound andcompression throttle valves; configuring a throttle valve and a blow-offvalve in an active suspension system to facilitate communication betweenthem via fluid pressure; combining a throttle valve and a blow-off valvein a single diverter valve; use of leakage in a hydraulic motor tomitigate the effect of inertia; and shaping force/velocity curves of anactive suspension using passive valving.

Separate Rebound and Compression Throttle Valves

In active vehicle suspension systems comprising a hydraulic pump, it maybe desirable to configure such a system so that passive valving can beschematically placed in parallel or in series with the a hydraulic pumpso as to prevent an overspeeding condition due to high hydraulic flowrates during high speed suspension events. Also it may be desirable touse passive valving, placed schematically in parallel or in series withthe hydraulic pump, to limit, and/or control the damping forcesgenerated by the high hydraulic flow rates during high speed suspensionevents.

Communication Via Pressure Between Throttle Valve and Blowoff Valve

In active vehicle suspension systems comprising passive valvingschematically placed in parallel or in series with a hydraulic pump, itmay be desirable to use a separate valve to limit the maximum speed atwhich the hydraulic pump rotates, regardless of hydraulic flow rate, andseparate valve that limits and/or controls the damping force at highhydraulic flow rates during high speed suspension events. Whereby, theactivation of the speed limiting valve actuates the damping forcecontrol valve by the pressure generated when the speed limiting valveactuates.

Combination of Throttle Valve and Blow-Off Valve in One Diverter Valvein a Shock Absorber

In active vehicle suspension systems comprising passive valvingschematically placed in parallel or in series with a hydraulic pump, itmay be desirable to use a common valve that limits the maximum speed atwhich the hydraulic pump rotates, regardless of hydraulic flow rate,whilst simultaneously limits and/or controls the damping force at highhydraulic flow rates during high speed suspension events.

Use of Leakage in a Hydraulic Motor to Facilitate Fluid By-Pass of theMotor and to Mitigate the Effect of Inertia

Shaping of Force/Velocity Curves of an Active Suspension Using PassiveValving

In active vehicle suspension systems passive valves schematically placedin parallel or in series with a hydraulic pump may be used to passivelycontrol the damping force velocity curves. These passive valves can beactuated to provide passive damping force control by either hydraulicpressure, or hydraulic flow rate. The damping force/velocity curve canby higher than the maximum active, or lower than the minimum active,damping force/velocity once activated. The passive dampingforce/velocity can be readily tuned to suit any particular application.

In FIG. 5A a regenerative active/semi-active damper 5-100 that comprisesa hydraulic regenerative, active/semi active damper valve 5-500, and amonotube damper assembly 5-300, comprising passive valving placedschematically in parallel and in series with the damper valve 5-500, isshown.

Blow-Off Valve

In the embodiment of FIG. 5A, a compression and rebound Blow-Off Valve(BOV) assembly 5-200 is shown in the piston head. The purpose the BOV'sis to limit the maximum pressure that can exist in both the reboundchamber 5-25 (of FIG. 5E) and the compression chamber 5-54 (of FIG. 5F).The BOVs may remain closed, (thus not allowing fluid to pass from therebound to compression chamber—and vice versa), until a preset pressureis achieved. Once this preset pressure is reached, the BOVs may open toallow fluid flow from either the rebound chamber to the compressionchamber, or vice versa, by-passing the damper valve 5-500 therebylimiting the pressure drop across the piston 5-7, and hence limiting thedamping force co-efficient. The BOV assembly 5-200 comprises a springbiased check valves for rebound and compression pressure limitation.

Referring to FIGS. 5B, 5C and 5D, the compression spring biased checkvalve comprises a compression spring perch 5-3 that is supported to thedamper piston rod 5-1. The compression spring perch also axially locatesthe rebound bump stop 5-2 and may transfer the rebound bump stop load tothe damper piston rod. The compression spring perch 5-3 supports thecompression check valve spring 5-4. The spring shown is of a compressiontype but may be of a different type such as Belleville or wave springetc. The compression spring 5-4 is loaded against a seal plate backupwasher 5-5. The seal plate backup washer 5-5 is loaded against a sealingwasher 5-6 and transfers the force from the compression check valvespring, to the sealing washer 5-6 via the seal plate backup washer 5-5.The scaling washer 5-6 is a flat disc type washer that has a flatsealing face 5-15 that may generate a hydraulic seal against the piston5-7. The compression sealing washer 5-6 and the piston 5-7 may be madefrom a hard material, such as steel for example, for durability so as tomaintain a good sealing surface after many operations.

The piston 5-7 is held concentric on the damper piston rod 5-2 and isfixed in axial rigid connection with the piston rod by means of therebound spring perch 5-10 which is rigidly connected to the damperpiston rod 5-2. The rebound spring perch 5-11 supports the rebound checkvalve spring 5-10. The spring shown is of the Belleville type but may beof a different type such as a compression or wave spring etc. Therebound spring 5-10 is loaded against a seal plate backup washer 5-9.The seal plate backup washer 5-8 is loaded against a rebound sealingwasher 5-8 and transfers the force from the rebound check valve spring5-10 to the sealing washer 5-8 via the seal plate backup washer 5-9. Theconstruction of the rebound sealing washer 5-8 is similar to that of thecompression sealing washer 5-6.

The piston 5-7 may contain compression flow passages 5-15 which allowfluid flow from the compression chamber 5-54 to the rebound chamber 5-25and rebound flow passages 5-14 that allow fluid flow from the reboundchamber 5-25 to the compression chamber 5-54. The compression sealingwasher 5-6 blocks off the compression flow passage 5-15 by virtue of thecompression check valve spring preload force. The compression flowpassage presents an exposed area on the compression sealing washer 5-6,and pressure in the compression flow passage acting over this area maypresent a force on the compression sealing washer, counteracting theforce compression check valve spring preload force. The compression flowpassage may remain closed off until the pressure in the compression flowpassage generates enough force to overcome that of the compression checkvalve spring preload force. Once this force is reached, the sealingwasher may move away from the piston 5-7 opening the compression flowpassage and allowing fluid flow from the compression chamber to therebound chamber, thereby limiting the pressure that can be generated inthe compression chamber.

The rebound sealing washer 5-8 blocks off the rebound flow passage 5-14,by virtue of the rebound check valve spring preload force and operatesin a similar manner to the compression BOV assembly as described above.

The spring perches 5-3 and 5-11 may set an initial spring preload, thispre-load can be set by changing the length of the spring perches or bythe use of shims (not shown), so that an accurate, repeatable andtunable (from damper to damper) pre-load can be set. It is important tokeep the mass of the sealing washers as low as possible, to reduce theeffects of inertia affecting the pre-load during high accelerationdamper effects. The sealing washers need to remain as rigid as possibleso as to maintain a good hydraulic seal with the piston. To this regard,the lightweight back up washers 5-5 and 5-9 are used in conjunction withthe sealing washers 5-6 and 5-8.

The spring perches 5-3 and 5-11 may also contain a seal washer travellimiter so as not to over compress the spring during high speed/highpressure damper events.

By varying the flow passage areas, the spring preloads, spring rates andmax allowable sealing washer openings, the damping force profile in bothcompression and rebound can be limited and tuned passively in parallelwith the valve.

Although the preferred embodiment is to use compression and rebound BOVit is possible to use only a compression or only a rebound BOV, and theinventive methods and systems are not limited in this regard.

Although the BOV assembly 5-200 is shown located in the piston head ofthe damper 5-100, this assembly can be located in other areas such asthe damper valve 5-500 itself.

Although FIG. 5-100 relates to a regenerative active/semi-active dampervalve in conjunction with a monotube damper configuration, the passivevalving described above can be used with a regenerative active/semiactive damper valve in conjunction with a twin tube or triple tubedamper configuration and the patent is not limited in this regard.

Although the preferred embodiment is to use compression and reboundthrottle valves in conjunction with compression and rebound BOVs, it ispossible to use any of these valves on their own or in any combinationof each other, and the patent is not limited in this regard. It is alsopossible to use a different type of spring biased check valve other thana seal plate configuration, such as a ball check valve, poppet checkvalve, or spool type check valve. These mechanisms are well known in theart and can be easily inserted in place of the seal plate check valve,and therefore the patent is not limited in this regard.

Throttle Valve

In the embodiment of FIG. 5A, a compression and rebound Throttle Valve(TV) assembly 5-400 and 5-300 respectively, is shown. The compression TV(CTV) is located in the top of the damper assembly 5-100 and the reboundTV (RTV) is located at the base of the damper above the damper floatingpiston.

Referring to FIGS. 5A & 5E; the rebound TV 5-300 comprises a RTV body5-16 that is located concentrically to the damper assembly 5-100, andlocates the damper pressure tube 5-24. The RTV body 5-16 locates the RTVseal block 5-17 and RTV seal washer 5-21. RTV preload washer 5-22 andRTV seal washer stop 5-20. The RTV seal washer 5-21 is held against theRTV seal washer stop 5-20 with a pre-load by springs 5-23 that pushagainst the preload washer 5-22. When the RTV seal washer is heldagainst the RTV seal washer stop, a circular flow passage 5-28 isgenerated between the RTV seal washer 5-21 and RTV seal block 5-17.There are flow passages 5-19 in the RTV seal block 5-17 that allow fluidcommunication between the rebound chamber 5-25 and the rebound side ofthe hydraulic pump/motor of the valve assembly 5-500, as shown by theflow arrows 5-27. Therefore, when the damper is in rebound, fluid flowsfrom the rebound chamber 5-25, through the circular flow passage 5-28and through the flow passages 5-19 in the RTV seal block 5-17, to therebound side of the hydraulic pump/motor of the valve assembly 5-500.The relatively small circular flow passage 5-28 offers a restriction tothis flow and may cause a pressure drop across the RTV seal washer 5-21that is proportional to the flow, this may generate a force imbalanceacross the RTV seal washer 5-21 counteracting the preload on the RTVseal washer from the springs 5-23. As the rebound flow increases thepressure drop and hence force imbalance across RTV seal washer 5-21 mayincrease until the force imbalance becomes greater than the springpreload, whereby the RTV seal washer 5-21 may start to close toward theRTV seal block 5-17. As the RTV seal washer 5-21 closes toward the RTVseal block 5-17, the circular flow passage 5-28 decreases in size andhence increases the pressure drop and force imbalance, thereby causingthe RTV seal washer 5-21 to close even further, until it becomes fullyclosed against the RTV seal block 5-17. The circular flow passage 5-28may now be completely closed as shown in FIG. 5H. The RTV is thereforeflow activated, and since rebound flow is proportional to rebound dampervelocity, the RTV is activated at by rebound damper velocity. Byadjusting the preload on the springs 5-23 and/or the size of thecircular flow passage 5-28, this velocity at which the valve activatescan be readily tuned.

When the RTV 5-300 is in the activated position, (as shown in FIG. 5H),flow to the rebound side of the hydraulic pump/motor of the valveassembly 5-500 is severely restricted, forcing fluid through smallorifices 5-30 in the RTV seal washer 5-21. This may cause the pressurein the rebound chamber 5-26, to sharply rise. When used in conjunctionwith a rebound BOV as described above, the pressure rise may cause therebound BOV to activate to limit the pressure in the rebound chamber5-26, and hence limit the rebound force. The orifices 5-30 in the RTVseal washer 5-21 are sized such that the flow rate to the rebound sideof the hydraulic pump/motor of the valve assembly 5-500, before the RTVseal washer 5-21 is activated, is similar to that of the flow rate tothe rebound side of the hydraulic pump/motor of the valve assembly5-500, after the RTV seal washer 5-21 activated, when flow is forcedthrough the orifices 5-30 at the higher pressure.

With this TV and BOV arrangement, flow to the rebound side of thehydraulic pump/motor of the valve assembly 5-500 may become restrictedat a predetermined rebound damper velocity, thereby restricting therotational speed of the hydraulic pump/motor at speeds above theactivation velocity. This may protect the hydraulic pump/motor againstoverspeeding when high damper velocities are achieved, and the dampingcoefficient may be determined by the passive BOV. As mentionedpreviously, the activation point and damping curve can be tuned to suita given application. This may provide a damping force curve similar tothat shown in FIG. 5V, where the curve from zero to the point 5-89, isthe maximum damping force generated by the hydraulic regenerative,active/semi active damper valve 5-500, and the point 5-90 is where boththe TV and BOV are fully activated, and the curve after the point 5-90is the passive damping force generated by the BOV flow characteristics.

It is possible to provide a smooth transition between points 5-89 and5-90 by controlling how the BOV assemblies open. One method to achievethis may be to replace the one piece sealing washer 5-6 and 5-8 with astack of flex washers that can vary the opening to the flow passages5-14 and 5-15 due to flexure of the flex washer stack under pressure.Another method to achieve a smooth transition is to use a series ofsealing washers that may provide staggered opening arrangement. Thesemethods are well known in the art and the patent is not limited in thisregard.

Referring to FIGS. 5F & 5G, the compression TV (CTV) 5-400 operates in asimilar manner to that of the RTV 5-300, and operates in conjunctionwith the compression BOV to achieve similar high-speed protection andpassive damping coefficients.

In the embodiment shown in FIGS. 5J, 5K & 5L, the compression andrebound TV are co-located at the top of the damper, (i.e. at the reboundend). In this embodiment, the TV sealing washer 5-21 is held in acentral location between the rebound and compression seal blocks 5-17and 5-17B respectively, by preload washers 5-22 and 5-22B, and springs5-22 and 5-22B, so that the rebound circular flow passage 5-28 issimilar to that of the compression circular flow passage 5-28B.

When the direction of flow is from the rebound chamber to the hydraulicpump/motor of the valve assembly 5-500, the TV may activate to block offlow to the hydraulic pump/motor of the valve assembly 5-500 from therebound chamber, as shown in FIG. 5L.

When the direction of flow is from the compression chamber to thehydraulic pump/motor of the valve assembly 5-500, the TV may activate toblock of flow from the hydraulic pump/motor of the valve assembly 5-500from the rebound chamber, as shown in FIG. 5 k.

The operation of the TV shown in FIG. 5J and activation set point issimilar to that described for FIG. 5E.

Diverter Valve

In the embodiment of FIG. 5A, a compression and rebound Throttle Valveassembly 5-400 and 5-300 respectively, working in conjunction with acompression and rebound BOV assembly is disclosed. As describedpreviously, the TV assembly may limit the flow (and hence speed) of thehydraulic valve 5-500, once a certain damper velocity is achieved, andmay generate pressures high enough to activate the BOV, thereby reachingthe maximum damper force. Even though this system may allow the damperto reach high velocities (approx. above 1 m/s for passenger vehicleapplications), without overspeeding or damaging the hydraulic valveassembly 5-500, this arrangement forces the damper curve to go sharplyto its maximum value once the TV activation damper velocity is reached.This may not offer the best high speed damping coefficients for certainapplications, whereby a tunable intermediate passive damping coefficientis desirable. To achieve a tunable intermediate passive dampingcoefficient, whilst providing high-speed protection for the hydraulicvalve, a diverter valve (DV) assembly, as shown in FIG. 5M, isdisclosed.

Referring to FIG. 5X a regenerative active/semi active damper 5-600 thatcomprises a hydraulic regenerative, active/semi-active damper valve5-500, and a monotube damper assembly 5-300, comprising a rebounddiverter valve (RDV) 5-800, and a compression diverter valve (CDV) 5-900is shown.

Referring to FIGS. 5M, 5O, 5Q & 5S the RDV 5-800 comprises a throttlebody 5-46, a sealing washer 5-50 and a seal body 5-47. The seal body5-47 is held concentric to the damper body of 5-600 and locates thedamper pressure tube 5-24. The seal body 5-47 also locates and seals offa middle tube 5-48. This may provide a first annular flow passage 5-51,between the pressure tube and middle tube, that is in fluidcommunication with the first port of the hydraulic pump/motor of thehydraulic valve 5-500, via a connector tube 5-58. A second annular flowpassage 5-53, is generated between the middle tube 5-48 and the damperbody of 5-600 that is in fluid connection to the second port of thehydraulic pump/motor of the hydraulic valve 5-500.

The seal washer 5-50 is held against the seal body 5-47 by springs 5-23.(shown in FIG. 5S), so that a relatively small circular flow passage5-49 is generated between the seal washer 5-50 and throttle body 5-46.The seal block also may contain small flow orifices 5-69 that are influid communication with the first annular passage 5-51, and when theseal washer 5-50 is held against the seal body 5-47 by springs 5-23, theseal washer 5-50 blocks off the small flow orifices 5-69, so that noflow exists between the rebound chamber 5-26 and the first annularpassage 5-51.

There are flow passages 5-52 in the throttle body 5-46 that is in fluidcommunication with the second annular flow passage 5-53, and hence thesecond port of the hydraulic pump/motor of the hydraulic valve 5-500.The rebound chamber 5-25 is in fluid communication with the circularflow passage 5-49, and the flow passages 5-52 in the throttle body 5-46,as shown by the flow arrows, 5-57. Therefore, when the damper is inrebound, fluid flows from the rebound chamber 5-25, through the circularflow passage 5-49, through the flow passages 5-52 in the throttle body5-46, and to the second port of the hydraulic pump/motor of thehydraulic valve 5-500, via the second annular flow passage 5-53, asshown by flow arrows 5-26 and 5-55. The relatively small circular flowpassage 5-49 offers a restriction to this flow, and may cause a pressuredrop across the seal washer 5-50 that is proportional to the flow, thismay generate a force imbalance across the seal washer 5-50,counteracting the preload on the seal washer from the springs 5-23. Asthe rebound flow increases the pressure drop, and hence the forceimbalance across seal washer 5-50, may increase, until the forceimbalance becomes greater than the spring preload whereby, the sealwasher 5-50 may start to close toward the throttle body 5-46. As theseal washer 5-50 closes toward the throttle body 5-46, the circular flowpassage 5-49 decreases in size and hence increases the pressure drop andthe force imbalance thereby, causing the seal washer 5-50 to close evenfurther, until it becomes fully closed against the throttle body 5-46.The circular flow passage 5-49 may now be completely closed, as shown inFIG. 5O. The RDV is therefore flow activated, and since rebound flow isproportional to rebound damper velocity, the RDV is activated at byrebound damper velocity. By adjusting the preload on the springs 5-23and/or the size of the circular flow passage 5-49, the velocity at whichthe valve activates can be readily tuned.

When the RDV 5-800 is in the activated position, (as shown in FIG. 5O),flow to the second port of the hydraulic pump/motor of the valveassembly 5-500 is severely restricted, forcing fluid through smallorifices 5-56 in the seal washer 5-50, as shown by flow arrows 5-57.This may limit the speed at which the pump/motor of the assembly 5-500rotates when the RDV is activated.

As the seal washer 5-49 closes toward the throttle body 5-46, it movesaway from the seal body 5-47, opening the small flow orifices 5-69 thatare in fluid communication with the first annular passage 5-51. This maynow allow fluid flow from the rebound chamber 5-26 to the first annularpassage 5-51, via the small flow orifices 5-69. As well as being influid communication the second port of the pump/motor of the hydraulicvalve 5-500, the first annular passage 5-51 is also in fluidcommunication with the compression chamber 5-24, via flow passages 5-62in the CDV throttle body 5-60, as shown in FIG. 5N.

Therefore, when the RDV 5-800 is activated, it may allow flow from therebound chamber 5-26 to two distinct flow paths: the first flow path isto the second port of the pump/motor of the hydraulic valve 5-500, viathe orifices 5-56 in the seal washer 5-50, and the second flow path isto compression chamber, via the first annular passage 5-51, and flowpassages 5-62 in the CDV throttle body 5-60. In this manner, the RDV5-800 diverts flow from the primary flow path—the second port of thepump/motor of the hydraulic valve 5-500, to a secondary flow path—thecompression chamber 5-24. This has the effect of limiting flow to thepump/motor of the hydraulic valve 5-500, whilst short-circuiting flowfrom the rebound chamber 5-25 to the compression chamber 5-24.

Since the flow to the compression chamber 5-24 is via the small floworifices 5-69 in the seal body 5-47, the pressure/flow characteristic ofthis flow path can be readily controlled to provide the desired passivedamping coefficient when the damper velocity is at a high enough speedto activate the diverter valve. As well as varying the orifice flowcoefficient, the distance that the seal washer 5-50 moves away from theseal body 5-47 can be varied to vary the flow coefficient. Also, theseal washer 5-50 may constructed of a stack of flex washers (as opposedto one, stiffer, washer) that can vary the opening to the small floworifices 5-69, due to flexure of the flex washer stack under increasingpressure in the rebound chamber. These types of valves are well known inthe art and the patent is not limited in this regard. Due to theflexibility of how the passive damper coefficient can be tuned, thepassive damper coefficient can be higher than the maximum damper forcegenerated by the hydraulic regenerative, active/semi active damper valve5-500, or lower than the minimum damper force generated by the hydraulicregenerative, active/semi-active damper valve 5-500, or anywhere inbetween, as shown in FIG. 5W.

When the seal washer 5-50 is held against the seal body 5-47 by springs5-23, the small flow orifices 5-69 in the seal body 5-47 present an areaon the seal washer 5-50, and any pressure differential that existsbetween the first annular passage 5-51 and the second annular passage5-53 (due to the pressure differential between the rebound andcompression chambers due to the damper force), may generate a force onthe seal washer due to the area presented on the seal washer. This forcemay act in parallel to the force imbalance on the seal washer 5-50 thatis generated by the fluid flow through the small circular flow passage5-49 that causes the RDV to activate. Therefore, by controlling thepressure differential between the first annular passage 5-51 and thesecond annular passage 5-53, the force imbalance, and hence theactivation point, on the RDV can be controlled. Since the differentialbetween the first annular passage 5-51 and the second annular passage5-53 is controlled by the hydraulic regenerative, active/semi activedamper valve 5-500, the damper velocity at which the RDV activates cannow be controlled by varying the damper force via the hydraulicregenerative, active/semi active damper valve 5-500. The loading on thehydraulic regenerative, active/semi active damper valve. 5-500 can beaccurately controlled so as to smooth out the transition to passivedamping when the RDV activates, thereby improving the ride quality ofthe damper.

Since the passive damper coefficient after the RDV has been activatedcan be readily tuned to be either greater or lower than the maximumdamper force, and the damper velocity at which the RDV activates can becontrolled by the hydraulic regenerative, active/semi active dampervalve, a broad damper force curve, similar to that shown in FIG. 5W canbe achieved, whereby, the activation velocity at max damper force isshown by point 5-81, the activation velocity at min damper force isshown by point 5-85, and the curve 5-83 represents the maximum tunedpassive damping coefficient after the RDV has activated, and the curve5-85 represents the minimum tuned passive damping coefficient after theRDV has activated. The area 5-85 between the maximum and minimum tunedpassive damping coefficient curves 5-83 and 5-84 respectively, is thebroad range to which the passive damping coefficient can be tuned, tosuit any particular application.

When the damper is in compression, fluid may flow from the second portof the hydraulic pump/motor of the hydraulic valve 5-500, through thesecond annular flow passage 5-53 into the rebound chamber 5-26. Fluidmay be in communication from the compression chamber 5-54 to the firstannular passage 5-51, via the CDV 5-900. The pressure in the compressionchamber 5-54 may be proportional to the compression damping force, andthis pressure may be present at the small flow orifices 5-69. Due to thearea exposed on the seal washer 5-50 from the small flow orifices 5-69,the compression chamber pressure may generate a separating force on theseal washer, counter-acting the preload placed on the seal washer 5-50from the springs 5-23. Once the separating force becomes greater thanthe preload force, the seal washer 5-50 may start to move away from theseal body 5-47, allowing fluid to flow from the first annular passage5-51 (and hence the compression chamber 5-54) to the rebound chamber5-25. This may limit the pressure that can be achieved in thecompression chamber, and thereby the RDV may now act as a compressionBOV, when the damper is in compression. Although the diverter valveoffers blow-off functionality, it might be desirable to use another BOVin acting with, or instead of, the diverter valve BOV. This other BOVcould be in the form of the embodiment of FIG. 5B, for example, and thepatent is not limited in this regard.

Referring to FIGS. 5N, 5P, 5R & 5T; the compression diverter valve (CDV)5-900 operates in a similar manner to that of the RDV 5-800, andoperates to limit the pump/motor speed of the hydraulic valve 5-500 whenthe damper is at high compression damper velocities, and to provide abroad passive compression damper coefficient after the CDV has beenactivated, as well as to act as a rebound BOV limiting the maximumrebound pressure when the damper is in rebound.

Although the damper architecture shown in the above figures is that of amonotube arrangement, the valving described above can be used in ahydraulic regenerative, active/semi active damper valve that isincorporated in a twin tube or triple tube damper architecture, and thepatent is not limited in this regard.

A device to improve the high-speed control of a hydraulic regenerative,active/semi active damper valve, and provide tunable high velocitypassive damping coefficients.

In order to provide active/semi active damping authority with reasonablesized electric motor/generator and hydraulic pump/motor, a high motionratio is required between damper velocity and motor rpm. Although thismay allow for accurate control of the damper at low to medium dampervelocities, this ratio can cause overly high motor speeds andunacceptably high damping forces at high velocity damper inputs. (referFIG. 5U). To avoid this, passive valving can be used in parallel and inseries with the hydraulic regenerative active/semi active damper valve.

The embodiments disclosed provide an economical, repeatable and tunablepassive damping force co-efficient that operates in concert with theregenerative, active/semi active damper valve.

Smart Valve—Coupling ECU, Motor, Hydraulic Pump and Sensor into One Unit

An active suspension system may be embodied as a complete unit that mayonly require configuring the unit in place of a standard shock absorberand providing power to it. Such a complete unit may be referred toherein and elsewhere as a smart valve. A smart valve may include anelectronic control unit or controller, an electric motor, a hydraulicpump, and one or more sensors configured into a single unit. In generala smart valve is able to create controlled forces in multiple (e.g.,typically three or four) quadrants of a vehicle suspension forcevelocity curve. Various embodiments of a smart valve are possible andmay optionally include the items identified above along with a pistonactivated hydraulic actuator, and the like. Constraints on a size and/orshape of a smart valve may indicate that a smart valve may need tocomply with sizing and shape/form factor constraints of passivedamper-based suspension systems, such that, among other things, a smartvalve based actuator may be installed in existing vehicle platformswithout requiring substantial re-design of such platforms.

A smart valve may comprise the components mentioned above (controller,electric motor, hydraulic pump, and one or more sensors) and may operateby the controller controlling the speed of the electric motor byapplying a current through the motor windings, effectively presenting atorque that resists the rotation of the motor. Such a smart valve may behoused in a single body. Alternatively the controller, electric motor,and sensors may be housed in a housing that can be assembled to ahousing for the hydraulic pump to facilitate communication among theactive suspension system components.

Configurations of a smart valve may include an electric motor, electricmotor controller and sensor in a single housing. Another configurationof a smart valve may include an electric motor, electric motorcontroller, hydraulic pump in a housing. In a variation of thisconfiguration, the housing may be fluid filled. An alternateconfiguration of a smart valve may include a hydraulic pump, an electricmotor that controls operation of the hydraulic pump, an electric motorcontroller, and one or more sensors in a single body housing. In yetanother configuration of a smart valve, the smart valve may include anelectric motor, a hydraulic pump, and a piston equipped hydraulicactuator that facilitates communication of hydraulic actuator fluidthrough a body of the actuator with the hydraulic pump.

Other configurations of a smart valve may be constrained by dimensionsand orientation that are dictated by a vehicle wheel well. Therefore asmart valve that is compatible with a wheel well may include a pistonrod disposed in an actuator body, a hydraulic motor, an electric motor,an electric controller for controlling the electric motor, and one ormore passive valves disposed in the actuator body, wherein the passivevalves operate in either series or parallel with the hydraulic motor.

Smart valve active suspension systems may be configured so that controlelectronics that provide currents to control the electric motor may beclosely integrated with the goal of minimizing the length of a highcurrent path from the control electronics to the electric motor.

While these and other smart valve configurations are possible, it may bedesired to configure a vehicle active suspension system that controlsall wheels of the vehicle. Such a system may include a plurality ofsmart valves, each being disposed proximal to a vehicle wheel so thateach smart valve produces wheel-specific variable flow and/or variablepressure of fluid in the valve by controlling an electric motor that isdisposed proximal to a hydraulic motor (e.g. coaxially) for controllingmovement via the hydraulic motor.

A smart valve may be configured in a variety of other ways. Someexemplary ways may include: the electronic motor controller isintegrated into the motor housing so that there are no exposed orflexing wires that carry the motor current to the motor controller; thesmart valve components are fully integrated into a damper body; thesmart valve components are integrated into a hydraulic shock absorberbody; the smart valve electronics may be mounted to the actuator; thehydraulic pump and electric motor of the smart valve are disposed on thesame shaft; the smart valve requires no hydraulic hoses; the hydraulicmotor is roughly axial with a piston rod of the actuator; the hydraulicmotor is roughly perpendicular to the piston rod travel direction; thesmart valve is mounted between the top of a strut and a lower controlarm of a vehicle wheel assembly; and the like.

Deployment environments of a smart valve may indicate certain size,shape, and orientation limitations. Exemplary smart valve embodimentsfor various deployment environments are now presented. An integratedsmart valve active suspension actuator occupies a volume and shape thatcan fit within a vehicle wheel well and between the damper top andbottom mounts. An integrated active suspension smart valve actuatoroccupies a volume and shape such that during full range of motion andarticulation of the damper, adequate clearance is maintained between thesmart valve and all surrounding components. An active suspensionactuator supports the smart valve co-axially with the damper body andconnects to the damper top mount. An active suspension actuator supportsthe smart valve co-axially with the damper body and occupies a diametersubstantially similar to that of an automotive damper top mount andspring perch. An active suspension pump and motor is configured to beless than 8 inches in diameter and 8 inches in depth, and even in somecases, substantially smaller than this footprint.

A smart valve may be self-contained and may not require externallygenerated knowledge, sensors, or other data from a vehicle. A smartvalve with integrated processor-based controller may performindependently, including self-calibrating regardless of whether thereare other smart valves (e.g. corner controllers) operating on otherwheels of the vehicle. A smart valve may deliver a wide range ofsuspension performance from operating as a passive damper, a semi-activesuspension, a variable suspension, a fully active suspension, and thelike because it is self-contained. A self-contained smart valve may becombined with a wide range of advanced vehicle capabilities to deliverpotentially more value and/or improved performance. Combining a smartvalve with predictive control, GPS enabled road condition information,radar, look-ahead sensors, and the like may be readily accomplishedthrough use of a vehicle communication bus, such as a CAN bus.Algorithms in the smart valve may incorporate this additionalinformation to adjust suspension operation, performance, and the like.In an example, if a smart valve of a rear wheel had knowledge of actionsbeing taken by a front wheel smart valve and some knowledge of vehiclespeed, the suspension system of the rear wheel could be prepared torespond to a wheel event before the wheel experiences the event.

Therefore, methods and systems of active suspension may include a smartvalve that may comprise an electric motor, electric motor control, oneor more sensors, a hydraulic pump, an actuator, and fluid tubes in aconsolidated enclosure that is compatible with vehicle wheel wellclearance constraints.

Referring to FIG. 6A, an active suspension actuator 6-100 with anintegrated Smart Valve 6-200 with electronics on the back of theelectric motor.

Referring to FIG. 6B, an active suspension actuator 6-400 with anintegrated Smart Valve 6-500 with electronics on the side of theelectric motor and valve housing.

Referring to FIG. 6C, the preferred embodiment of electro-hydraulicregenerative/active smart valve 6-100 comprises of an electro-hydraulicvalve assembly 6-1 coupled with the controller module 6-22, wherein thecontroller module is situated on the top of the electro-hydraulicregenerative/active smart valve 6-100.

The electro-hydraulic valve assembly 6-1 comprises a hydraulicpump/motor assembly 6-2 close coupled to the rotor 6-4 of an electricmotor/generator, wherein the stator 6-3 of the electric motor/generatoris rigidly located to the body of the electro-hydraulic valve assembly6-1. A rotary Hall effect position sensor 6-13, that measures therotational position of a source magnet 6-15 that is drivingly connectedto the electric motor/generator rotor 6-4 is mounted directly to thelogic subassembly 6-221 x and is protected from the working hydraulicfluid of the electro-hydraulic valve assembly 6-1 by a sensor shield6-14.

In the embodiment of FIG. 6C to achieve the most efficient heatdissipating capability the power pack unit 6-223 x may be a) mountedflat on the surface of the valve assembly 6-1 with all six FETs or IGBTsutilizing this surface as a heat sink.

The controller module 6-22 comprises the logic subassembly 6-221 x,power storage unit 6-222 x and energy dissipating power pack 6-223 x. Apower storage unit 6-222 x is mounted on the side of the hydraulic valveassembly 6-1 or can be integrated with the power pack 6-223 x

The embodiment of FIG. 6D is similar to that of the embodiment of FIG.6C, with the power pack 6-223 x mounted to a dedicated heat sink that isthermally decoupled from the hydraulic valve assembly 6-1. A powerstorage unit 6-222 x is mounted on the side of the hydraulic valveassembly 6-1 or can be integrated with the power pack 6-223 x

Yet in another preferred embodiment the power pack 6-223 x is split intothree subunits 6-2231 a(b,c) with each subunit housing a single leg(half bridge) of the power pack. It is well understood to any skilled inthe art that many other permutations of the preferred embodiments can beimplemented within the scope of the current inventive methods andsystems.

For the purpose of minimizing thermal load and volume the logicsubassembly 6-221 x may be subdivided into logic power module, sensorinterface module and a processor module. In one preferred embodiment thelogic subassembly 6-221 x may be using a rotor position sensor 6-13. Therotor position sensor may share the same pcb that is used for housingFETs (IGBTs) or may be mounted on a flex cable. In another preferredembodiment the logic subassembly 6-221 x may be completely sensorless.

Furthermore, it is well understood that all the components of thecontroller module 6-22 can be integrated into a single assembly andproduced on a single PCB.

FIG. 6E demonstrates a generic Smart Valve architecture. The Smart Valve6-40 may contain an electronic controller 6-41, which may contain aprocessor 6-44 and a motor controller. 6-45, along with sensor andcommunications inputs and outputs 6-43, 6-42. The controller isconnected to the terminals of the electric motor 6-46 via electricalwires or a direct connection, such as three phase wires. The shaft ofthe motor is operatively coupled to the shaft of a hydraulic pump thatmay be both bidirectional and backdrivable (but may also beunidirectional or pumping only).

FIG. 6F demonstrates an integrated active suspension system thatpackages into the wheel-well of a vehicle, substantially between theupper 6-54 and lower 6-55 suspension members. The actuator body 6-51stands in the place where a typical damper does. This is typicallyconnected between the upper top mount and the lower control arm,depending on suspension configuration. A motor 6-46 and hydraulic pump6-47 are close-coupled to the actuator body 6-51 in the wheel well 6-56,and in many cases substantially between the suspension members. One ormore passive valves 6-52 may also be integrated into the damper body.

FIG. 6G demonstrates an embodiment of an inertia cancelling controlsystem for an active suspension. A torque command 6-57 typically from aseparate control loop is added to the inertial cancellation torquecommand 6-58 from the last time step. This results in a total torquerequest to the motor controller 6-59, which results in rotaryacceleration 6-60 (which may be zero). The rotary acceleration 6-60 ismultiplied 6-61 by the total rotary moment of inertia of the system 6-62to obtain the desired inertia cancellation torque 6-63. This is fed intoa filter 6-64, which may provide compensation for system dynamics, phaseloss, delays, and high frequency noise, among other tasks. The output ofthe filter 6-58 is then fed back into the positive feedback loop at thenext time step.

FIG. 6H demonstrates an embodiment of the control signals of a torqueripple cancelling control system. Without the active ripplecancellation, a torque command 6-67 results in a pressure ripple in thehydraulic system 6-68 about the desired pressure 6-69. This has severalundesirable effects such as audible noise. With the active ripplecancellation, the torque command 6-70 is added to a signal which modelsthe predicted ripple and inverts it 6-71. The result of this controlsystem is a fairly flat pressure signal 6-72 for a flat torque curve.

Sensor Algorithms

A processor controlled active suspension system may include sensors fordetecting various conditions that can be used to improve or optimizesuspension performance. As noted elsewhere herein, an active suspensionsystem that includes a 3-phase electric generator for controllingsuspension operation may require accurately detecting a position and anacceleration of the electric generator rotor. Algorithms may be appliedto rotor position sensed data to deliver a highly accurate positiondetermination. Algorithms for self-calibration of rotor position sensingmay include: detecting noise patterns in the data that are filtered outby selective position sensing; real-time on-line latency-free rotationalsensor calibration based on an off-line generated calibrationcurve/model; determining a position error and therefore a compensationfunction based on velocity calculation and velocity ratio calculation.Calculating position of the rotor position accurately may be combinedwith controlling motor windings current values to develop a rotationalposition coordinate system relative to a rotor-axis of a 3-phasebrushless motor. These and other sensor-related algorithms for a vehiclesuspension system are described further below herein.

Electric motor controls rely on knowledge of the position of the rotorwith respect to the stator at any time in order to correctly align thephase of the rotating magnetic field with respect to the stationarymagnetic field. Especially for applications involving low-speed and hightorque operation, where sensorless techniques cannot be used, a positionsensor is required and the cost of this sensor can be of significantimpact on the system design.

A low quality sensor reading can introduce large errors, especially whenthe sensor output is used to derive calculated quantities, such asvelocity and acceleration. Lower cost sensors in general tend to exhibitmore pronounced output errors. These errors can be of many differentvarieties, but can be grouped into major functional groups:

1) Errors that exhibit no correlation with the sensor reading or othereasily measurable external factors, such as electrical noise,discretization or quantization errors, or similar

2) Errors that correlate with external influences, such as temperatureerrors, pressure errors, humidity errors, or similar

3) Errors that exhibit correlation with the sensor reading, such ascalibration errors, position-dependent errors, velocity-dependenterrors, or similar

In this document, we may focus on the third type of errors, whichcontain a repeated pattern over the range of operation of the sensor.FIG. 7C explains the relationship between actual measured quantity (inthis case position) and the output of a typical sensor with errors ofthis type. Curve 7-14 shows the ideal output for a typical sensor, whichperfectly follows the measured quantity across its full range. Curve7-15 on the other hand shows a typical output signal with somerepeatable deviation from the measured quantity over the range ofoperation of the sensor.

Inventive Methods and Systems

The present inventive methods and systems allows for calibration of alow quality sensor to produce a low-latency, high accuracy outputsignal.

In one embodiment, this inventive methods and systems can be applied toa position sensor in a rotary three-phase brushless electric motor. Thesensor is a low-cost, low-resolution magnetic rotary position encoderthat exhibits strong deviation of the measurement from the actualposition in part due to sensor misalignment, sensor assembly errors, andmaterials tolerances.

The measured position signal can be decomposed into the actual position,an error that is strongly correlated with the actual signal, plus anyerror not correlated with the output signal. This can be written in theformP _(measured) =P _(actual) +e _(c)(P _(actual))+e _(u)   EQUATION 1

Where P_(measured) is the output of the sensor, P_(actual) is the signalthe sensor is trying to read, e_(c) is the part of the error in thesensor output signal which is correlated with the actual measuredquantity (and is thus for this particular example a function of theactual position), and e_(u) is the part of the error in the sensoroutput signal which is uncorrelated with the actual measure quantity.

FIG. 7A shows a flow diagram for the process described here. Theposition signal Pan is differentiated in block 7-1 to create themeasured velocity, for example by using discrete-time differentiationalgorithms. The resulting signal can be written in mathematical form byderiving equation 1 to obtain:

$\begin{matrix}{V_{measured} = {\frac{\partial\left( P_{measured} \right)}{\partial t} = {{\frac{\partial\left( P_{actual} \right)}{\partial t} + {\frac{\partial\left( {e_{c}\left( P_{actual} \right)} \right)}{\partial P_{actual}}\frac{\partial\left( P_{actual} \right)}{\partial t}} + \frac{\partial e_{u}}{\partial t}} \approx {V_{actual} + {\frac{\Delta\;{e_{c}(P)}}{\Delta\; P}V_{actual}} + \frac{\partial e_{u}}{\partial t}}}}} & {{EQUATION}\mspace{14mu} 2}\end{matrix}$

The next step is to use a filter represented by block 7-2 in FIG. 7A toremove any component of this signal that is periodic with the position.In one embodiment, this filter has the shape shown in FIG. 7F. Like anyfilter, it exhibits some group delay, which must be taken into accountin the following steps.

Turning back to equation 2 and writing a filtered version of the outputsignal, we find that the error term correlated with position (Δe_(c) inequation 2) is canceled out by the filter and the remaining signal isapproximately delayed by the group delay in the filter. This leads tothe expression in equation 3

$\begin{matrix}{{V_{{measured},\;{filtered}} \approx {V_{{actual},\;{delayed}} + \left( \frac{\partial e_{u}}{\partial t} \right)_{filtered}}} = {V_{{actual},\;{delayed}} + {noise}}} & {{EQUATION}\mspace{14mu} 3}\end{matrix}$

We find that the result is a delayed estimate of the actual velocity,along with a “noise” term that represents any error uncorrelated to theposition signal.

We now use a time delay, described by block 7-5 in FIG. 7A, with amagnitude equivalent to the approximate group delay in filter 7-2, tocreate a delayed version of the measured velocity. This can be writtenas:

$\begin{matrix}{V_{{measured},\;{delayed}} = {{V_{{actual},\;{delayed}}\left( {1 + \frac{\Delta\;{e_{c}(P)}}{\Delta\; P}} \right)} + {noise}}} & {{EQUATION}\mspace{14mu} 4}\end{matrix}$

The next step in the process is to divide this signal byV_(measured, filtered), as represented by divider block 7-3 in FIG. 7A.Writing out the resulting equation, we see that we can estimate theincremental error at any given position from this calculation:

$\begin{matrix}{{\Delta\;{e_{c}(P)}} = {{\Delta\;{P\left( {\frac{V_{{measured},\;{delayed}}}{V_{{measured},\;{filtered}}} - 1} \right)}} + {noise}}} & {{EQUATION}\mspace{14mu} 5}\end{matrix}$

This error is now stored in a buffer 7-4. Any entry in this table at agiven position is then averaged over time in order to remove the effectsof any uncorrelated error signal. After only a few averages, the bufferthen may contain a very good estimate of the actual calibration error asa function of the measured signal.

The entire calculation is run in an asynchronous way, meaning the outputof the calculation does not affect the sensor reading at the presenttime step. Instead, once the buffer 7-4 may contain enough averages, thecorrection is simply added at each time step to the measured signal,thus removing any latency that would be present if we simply filteredthe signal through a time-based filter at any step. Also, by averagingthe correction over many cycles, we remove any uncorrelated error fromit, which would be impossible with simple filtering.

The signal is then used in an adaptive encoder model, such as the oneshown in FIG. 7B as one possible implementation. In this model, theparameters 7-7, the measured quantities 7-8, the estimated quantities7-9, and the corrected signal 7-10 along with calculated values 7-11 arefed into a model of the system.

In one example, the adaptive encoder model can then be used to correctthe readings from the sensor in case there are sensor offsets, drifts,or counting errors. The model estimates the voltage generated throughelectro-magnetic force (“back EMF” in industry parlance) from the rotorvelocity and the motor's voltage constant parameter. It can alsoestimate the phase angle of this voltage with respect to the threephases of the motor, given the position of the rotor with respect to thephases in the stator.

For the purposes of explaining the following figures, it is important tonote that for the motor control logic in this implementation adirect-quadrature (“DQ0”) control is used. This type of control is awell-known method in the industry, whereby the position of the movingcomponent of a3-phase motor (the “rotor” in a rotary motor, or“armature” in a linear motor) with respect to the stationary componentof the motor (the “stator”) is used to transform the system coordinatesinto a reference frame that rotates in the coordinates of the electricalphase angle, by using what is known in industry as the DQ0transformation. For further discussion purposes, this rotating referenceframe may forthwith be used as a basis.

FIG. 7E shows how the back EMF can be represented in DQ0 space by avector 7-30. This vector is defined by a characteristic angle 7-31 withrespect to the DQ0 reference frame (which rotates with respect to theelectrical phase angle as the rotor moves with respect to the stator),and by a component 7-32 aligned with the quadrature axis.

We start by assuming that the position signal is close to correct, andcalculate the total expected back EMF 7-30 and the measured back EMFcomponent 7-33 projected onto the D axis. This allows us to estimate theangle 7-31. That angle may be zero and the back EMF on the D axis may be0 if there is no error in the position signal.

The assumption is that any differences between model and measured signalmust first be due to sensor offsets, then to parameter errors whereconsistent, and last to sensor errors. Therefore, the model output canbe adjusted until it matches the measured response, thus finding thecorrect sensor offsets for example. This process must be done slowly inorder to react to overall trends but not to noise in the signal, butfast in order to avoid using false sensor information for too long. Inone embodiment, a simple low-pass filter can be used to trace both themeasured and model response, and once they deviate by more than acertain amount a correction can be applied.

FIG. 7D shows a possible embodiment of the complete motor controlstrategy for an active suspension system. A rotary position encodersignal (7-17) is read, then corrected by any necessary sensorcorrections due to offsets, drifts, or other errors as described above(signal 7-18 and block 7-26). The corrected signal is then fed throughthe calibration process described in FIG. 7A (block 7-19) to create ahigh quality position signal from which to derive velocity (signal 7-21)and acceleration estimates (signal 7-22).

Block 7-23 represents the DQ0 transformation of measured currents intodirect and quadrature currents, and the inverse transformation of thedirect and quadrature output voltages into the phase voltages.

From the desired torque 7-24 a desired current Iq is calculated, andthis signal is then compared to the measured currents in the DQ0 axes tocreate a control voltage output Vq,d. Finally, the control voltage istransformed by block 7-23 into the desired voltages at each switch ofthe 3-phase bridge (block 7-25), resulting in the pulse-width-modulatedswitch states for controlling the motor currents.

Vehicle Dynamics Algorithms

An active suspension system that includes corner controllers thatmonitor and control a hydraulic actuator may be configured with vehicledynamics algorithms to facilitate addressing situational active controlschemes, large event handling, power and energy management, and thelike.

Large event handling may include scenarios such as a speed bump that istaller than the operable stroke length of the active suspension systemactuator, traveling over a curb, very deep potholes, steep hills, andthe like. Algorithms for handling large events may take intoconsideration total suspension operable travel length, position withintravel length, change in position, rate of change in position, energyavailable for controlling the actuator, duration of event, and the like.A goal of such an algorithm may be to cause the actuator to lock outwith sufficient margin left in the operable travel length to avoiddamage to the actuator. A second goal may be to adapt the rate of travelas the actuator approaches its maximum operable travel length so as tomitigate the impact of the large event.

Situational active control schemes might include activating predictivealgorithms for a rear wheel action based on actual front wheel actions.When information about other wheels, particularly a leading wheel can begathered and analyzed, predictive algorithms may facilitate control ofthe actuator to further improve suspension performance. In an example,if a front wheel event that is caused by a speed bump is reported to therear trailing wheel, information such as total bump height, angle ofleading and trailing edges of the bump, and length of the bump may beused to configure real-time response algorithms that operate during anyevent to better handle the speed bump. In the example, if the bumpheight exceeds the actuator operable travel by twenty-percent of thebump height, it may be possible to adjust the real-time responsealgorithms to absorb eighty percent of the bump, thereby effectivelyreducing the impact to the vehicle to only twenty percent of the bumpheight, thereby making the vehicle movement over the bump smoother.

Energy management algorithms in an active suspension system may providediffering energy usage and regeneration models for various operatingmodes. Operating modes may include stopped mode, demo mode to showcontrol capabilities when the vehicle is stopped, road surface-specificmodes, driving condition-specific modes, modes for various vehiclestates (e.g. speed, acceleration, etc), event-based modes, combinationsof these and others.

Power and/or energy optimizing can benefit active damping of a vehiclewithout requiring input from or providing information to other cornercontrollers or vehicle systems. In general power and/or energyoptimizing algorithms may consider factors such as power limits, energylimits, power bus voltage, and the like to optimize power and/or energyusage while providing desired vehicle dynamics. Optimizing power and/orenergy usage may target an average power, peak power, or combinations ofboth depending on the vehicle operating mode. Vehicle dynamicsalgorithms may accept a measure of power and/or energy as an inputvariable. Such inputs may include average power, peak power, andcombinations of average and peak power. By considering a wide range ofvehicle dynamics-related factors as well as local corner controller andwheel related factors, it may be possible to throttle power demand/usagein a corner controller without critically disrupting vehicle dynamics.

Managing actuator performance based on power averaging in an activesuspension system may provide benefits to vehicle suspension operationover long time periods. When an active suspension system may closelycouple a processor-based corner controller with an electric generatorfor regenerating and controlling actuator operation, specificinformation about the amount of power being consumed or generated may beknown and therefore leveraged in power averaging algorithms. Actionsthat can be taken include, without limitation, throttling the peak powerthat the active suspension system can deliver and/or consume, and thelike. These peaks may also be dynamically moved based on a range ofconditions. As peak or average power limits are moved, depending on theroad, new values for these limits (e.g. 200 W per corner) can be set,such as to arrive at an average total power demand for the vehiclesuspension. In an example, when a suspension system is set to ‘sport’mode (stiffer suspension) monitoring power with an eye toward averaging,may prevent overly consuming stored energy. Overly consuming storedenergy (e.g. burning too much power in “sport” mode) may effectivelyresult in burning capacity that may be needed or may cause a net drainon energy sources, such as the vehicle electrical system. One solutionto avoid overly consuming stored energy is to establish a net powerconsumption average of zero. On a lengthy rough road, an activesuspension system may consume enough energy to require setting a zeroaverage for power consumption. The result may be degradation in activesuspension system performance due to a long continual power demand. On arelatively smooth road where the system regenerates a lot of energy,average power limits can be increased due to the gaining power.

A vehicle's suspension system has two main primary goals: to keep thedriver and passengers comfortable, and to maintain good contact betweenthe tires and the ground. These same goals remain true for an activesuspension, but due to the greater flexibility in the applied force, therequirements can be mapped out in a more detailed way as a function offrequency. FIG. 8B shows a possible way to describe the requirements andfrequency ranges in a typical modern automobile, mapped against thefrequency axis 8-8. While low-bandwidth active systems (8-9) control themotion of the car at up to maybe approximately 2 Hz, a higher bandwidthactive system (8-10) can control the motion of car and wheel up to 40 Hzand above. The ride quality challenges encountered at the lowestfrequencies are related to maintaining a good perception of connectionof the vehicle with the road; in the medium frequency range between 2and 8 Hz, driver and passenger comfort are key; between 8 and 15 Hz roadholding (including reducing tire force variation) is the main target,and in the higher frequency range above 15 Hz it is mainly harshness androad feel that drive the goals for the suspension.

An aspect of active suspension systems are their use of power highenergy consumption. In order to achieve the goals set above, the systemmust fight compliances and loss mechanisms inherent in the vehicle, suchas friction, suspension spring stiffness and roll bar stiffness,hydraulic losses, and damping in the various rubber elements (bushings),for a high percentage of its operating cycle. This leads to a largeconsumption of power in even the most efficient active systems. Byfocusing on the more important performance goals only, or by wateringdown performance in general, these systems can be made more efficient,though at the cost of significant reduction in the benefits the systembrings to the end consumer, as can be seen in some of the systems on themarket at the current time.

Inventive Methods and Systems

A better approach to solve this dilemma is what we call “situational”active control, whereby the amount of active control used is dictateddirectly by the situation at hand. FIG. 8A shows one such embodiment,where an “event detector” (block 8-1) reacts to inputs from sensors 8-2and estimates 8-3 to decide if an event requiring high amounts of activecontrol has happened, is in process, or is about to happen.

The sensor set used for this can include any of the many signalsavailable in a modern car, including acceleration sensors and rotationalrates of the car body (gyroscopes), position or velocity of thesuspension, vehicle speed, steering wheel position, and other sensorinformation such as look-ahead cameras. Estimated signals may includeestimated (current or upcoming) road vertical position, estimated roadroughness, position of the vehicle on the road, and other availablesignals.

For the rear wheels only, the information gathered from the frontwheels, such as estimated road position, input harshness, suspensiontravel history, or other useful signals, can then be used to improve theevent detection on the rear wheels. FIG. 8F shows a graphic of this.

The output of the event detector can be in the form of a command whenthe information is very good, or in the form of a parameter adjustment(such as a response to rough road or to driver input, where the responsemay be a change in the control strategy going forward), and may ingeneral be accompanied by a “confidence” factor. This output, along withvehicle feedback sensors 8-4 and measured driver input 8-5 is the inputto the actuator control logic 8-6, which determines the required outputcommand.

As an example, when the event detector recognizes an emergency maneuverthrough large lateral acceleration or longitudinal acceleration, itincreases the road holding ability and decreases the comfort in thesuspension.

This approach reduces the requirements for the active suspension, whichis essential in making it a reality in the automotive business. Only asystem that reduces its power output during times and driving scenarioswhere the consequent reduction in performance does not significantlyimpact the driver's and passengers' comfort or safety, and thus is notperceptible to the vehicle's occupants, can compete with much simpleralternatives existing on the market, which provide overall lessperformance but at a much smaller overall integration cost.

Power Averaging

An important aspect of any active suspension system is its powerconsumption. In order to allow the user to experience maximum comfort,while also maintaining average output or regenerated power at theexpected target levels, a scheme such as described in FIG. 8D is used.

The desired target power 8-24 is compared in the power averaging block8-25 to a calculated quantity correlated with the actual power output,calculated or measured, of the system. In one implementation, thiscalculated quantity is a filtered moving average of the power, thusproviding a low-noise representation of the mean power over the pastperiod of time. The difference between the two determines a variable wecall the power factor, which is used as input into the smart commandscaling block 8-27 along with the desired actuator command 8-26.

In one implementation, the actuator signal is then limited to a maximumallowable short-term output power, which is scaled by the power factor.Thus, at low power factors the suspension mostly regenerates power, thuslowering the average power output calculated in block 8-25. High powerfactors on the other hand allow the suspension to use as much power asneeded to achieve maximum comfort, and thus may often raise the averagepower output until the desired target value is reached and the powerfactor is lowered.

The command scaling can be done in many ways that allow for a goodcorrelation of power factor with average power output; these include butare not limited to limiting short or medium term output power in themotors, increasing short or medium term allowable regeneration, ormodifying the torque command consistent with other strategies forfinding a best possible approximation to the desired command whilereducing the power output such as for example reducing torque to itsnearest point to the equal power line.

In a different embodiment, the power factor can also be used to modifythe control gains inside the active feedback loop to increase its powerefficiency, for example by reducing the overall gain on the body control(which requires power during large part of its output) or by increasingg the gain on the wheel control (which I large part absorbs and thusregenerates power).

Hill/Bump Detection

An important aspect in an active suspension is the ability to predict ifa road feature is a “hill” or a “bump” FIG. 8C shows how we understandthe distinction between those two. Subplot 8-22 shows a series of bumps.A “bump” 8-17 is an oscillation in the road vertical displacement 8-20that does not exceed the full suspension travel 8-18 of the vehicle.This distinction is easy to understand if we assume the vehicle is kept“inertial” at all times, meaning that the vehicle glides and does notrespond to road input at all. In this case, as long as any road featuresare smaller than the suspension travel, then the control strategy mustremain the same.

Subplot 8-23 shows a “hill” 8-18, defined as a change in vertical roadposition that exceeds the full suspension travel 8-15. In this case, thesuspension control strategy must change in order to not reach the end oftravel of the actuator (which causes large discomfort in the vehicle andshould therefore be avoided as much as possible). Also, the transitionis smoother the earlier we can detect this event.

At point 8-21 the suspension system must look at the available sensorsand decide as soon as possible that this event is about to exceed thefull suspension travel. In absence of any sensors predicting the roadahead of the vehicle, this can best be done by using a combination ofthe road slope, vehicle attitude, driver input, and past history ofsensor signals.

The detection is based on pattern recognition: if the combination ofsensors follows an expected pattern, then the event is recognized as abump or hill. For example, if the road has been rough for the past shorthistory, and suspension motion has not exceeded a small percentage oftotal available travel, then we can assume the next input is also abump. Events with large slope in the road, especially if the driver ismaintaining elevated speed and not slowing down, also match the patternexpected for bumps.

Hills can be predicted through intermediate road slope (smaller thanthat of a typical bump, but high enough to be of concern) maintained fora length of distance greater than some lower threshold; also, driverinput factors such as a slowing down or a large steering input (as wouldbe typical for ramps and driveway entrances) can be used as leadingindicators.

Looking at the velocity at which the suspension is compressing orexpanding, and the distance left to travel in the direction it is movingin before reaching its limits, allows us also to predict the severity ofthe event, and thus how fast we must respond. An event is severe if theratio of closing velocity to distance left to target is high, and it isnot severe if that ratio is low compared to a threshold selected basedon driving data from typical roads. This threshold can also be adaptedaccording to the road type and past history of road profile.

An event with high pattern recognition, and high severity, causes animmediate reaction by the control system to both stiffen up the damperportion of the command, ease up on the skyhook portion of the command,and in the limit also command active force to prevent large impacts intothe travel limiters.

A system like this can use any sensor available in the vehicle or thesuspension. A minimum set may include position sensing on the damper andmaybe an acceleration signal on the wheel; in addition to that, vehiclespeed and steering angle are important, as is the in-plane motion of thevehicle to determine if the suspension motion is caused by the vehicleor the road. Any sensors able to “preview” the road, such as look-aheadcameras, laser signals, radar, or similar, can also be used to refinethe pattern recognition.

Model-Based Vehicle Control in an Active Suspension System

Active suspension systems are able to respond to road inputs in order toisolate the vehicle from the road, and to driver inputs in order todisconnect the vehicle's in-plane response (the acceleration, braking,and turning of the vehicle) from its response out of plane (thepitching, rolling, and heaving). In order to achieve high performance inthis, a typical control system must use high bandwidth sensors and closeloops around them, which can lead to high cost and can be difficult toachieve in the presence of external influences that modify the behaviorof the vehicle, aging of the vehicle, and similar structural changes tothe vehicle.

Inventive Methods and Systems

A different approach is presented in FIG. 8E. In order to best respondto the driver inputs 8-29, those inputs are measured by a sensor on thevehicle. These inputs include steering input, throttle or accelerationinput, brake input, and any other input commands by the driver.Typically in modern vehicles these inputs are already measured and usedfor other functions such as stability control, and thus can simply beread off the vehicle CAN or FlexRay bus. These sensor signals are theninput into a dynamic model representing the in-plane dynamics of thevehicle and the out-of-plane dynamics of the vehicle. A simplifiedrepresentation of the vehicle as a rigid structure can be very effectivefor this, for example a bicycle model as presented in many publicationswith in-plane and out-of-plane dynamics. Other parameters can also beentered into the model, such as the estimated road coefficient offriction (which is often calculated in anti-lock braking (ABS) systems),overall road conditions, and suspension roll angle with respect to theroad.

The outputs of the model 8-31 are the estimated states of the vehicle,most importantly the lateral acceleration component due to steeringinput, the fore-aft acceleration component, the estimated yaw rate, andthe vehicle's estimated sideslip angle. These estimates are then used tocalculate a required force command in block 8-35, based on knowledge ofthe vehicle's geometry and inertial properties. The output Force is thenadded to the feedback force coming from the closed-loop control system8-33, and the sum constitutes the total actuator input into the vehicle8-32. The benefit of this approach is that this component of the controlis almost completely open loop, meaning that the output of the open loopaction block 8-35 does not affect its input (the driver or road input8-29) in a strong way. Note that though the driver may react to actionsin the suspension through steering or braking, this reaction if presentat all may be, except for extreme cases, very small and slow and thusnot a concern from the vehicle dynamics point of view. An open loopcontrol system has the advantage that it can never cause instabilities,meaning that the control system's command outputs and resulting motionsof the vehicle may always remain within clear bounds and have apredictable behavior even when the model is completely inaccurate or thesensors are of low quality. The failure modes of such a system are thusvery easily handled and in general not dangerous, which is a greatadvantage for an application critical to occupant safety like avehicle's suspension system.

In order to improve the performance of this system, a Kalman filter canbe implemented, whereby the model parameters are adjusted using otheravailable sensors in the vehicle; for example, the longitudinalacceleration estimated from brake torque input can be compared to themeasured acceleration signal to improve the vehicle mass and the brakeperformance parameters used in the model.

Rear Wheel Predictor Model

FIG. 8G demonstrates a control system that improves controllability ofan adaptive suspension (either fully active or semi-active) by utilizinginformation from the front wheels in order to control the rear wheels.Here, a system is used that builds a complex three-dimensional model ofthe road with respect to the vehicle. The control system operates asfollows: a vehicle state estimator 8-43 uses sensor information such asaccelerometers, steering wheel angle, velocity and vehicle informationfrom third party sources (e.g. stability control system engagementinformation, ABS braking, brake position, etc.) in order to determinethe change in velocity over a time step of the vehicle in threedimensional space (lateral, longitudinal, vertical). This delta positioninformation is used to update a matrix 8-44 that may contain thevertical height (z position) of the road around the car. The matrix iscontinuously updated such that is holds a map of the height of the roadat positions around the vehicle, relative to the vehicle. The map isthen fed into a system and vehicle dynamics model 8-45, which usesinformation about upcoming road height to adapt damping levels, gains,and in some embodiments predicatively move the wheel.

While the present teachings have been described in conjunction withvarious embodiments and examples, it is not intended that the presentteachings be limited to such embodiments or examples. On the contrary,the present teachings encompass various alternatives, modifications, andequivalents, as will be appreciated by those of skill in the art.Accordingly, the foregoing description and drawings are by way ofexample only.

What is claimed is:
 1. A method of energy delivery to an activesuspension system, comprising: providing an active suspension system ina vehicle, the active suspension system comprising a hydraulic pump andan electric motor, the hydraulic pump and electric motor beingoperatively coupled, one or more sensors, an energy storage facility,one or more bypass valves, a controller, and one or more activesuspension actuators, each actuator comprising an actuator body, acompression volume, an extension volume, and a piston; detecting a roadinduced movement of a wheel of the vehicle with at least one of the oneor more sensors; sending data representing the movement from the atleast one of the one or more sensors to the controller; determining, atthe controller, based on the data representing the movement, a firsttorque command to be applied to a shaft of the hydraulic pump; receivinga required amount of energy at the electric motor from the energystorage facility; applying a first torque corresponding to the firsttorque command to the shaft of the hydraulic pump using at least aportion of the required amount of energy received thereby altering apressure in at least one of the compression volume or the extensionvolume; and modifying flow of hydraulic fluid to the hydraulic pump fromat least one of the compression volume or the extension volume with atleast one of the one or more bypass valves.
 2. The method of claim 1,further comprising flowing hydraulic fluid into and out of anaccumulator that is in fluid communication with at least one of thecompression volume and the extension volume.
 3. The method of claim 2,further comprising producing a second torque with the hydraulic pump dueto another road induced movement of the wheel, and driving the electricmotor as generator using the second torque to generate a second amountof energy.
 4. The method of claim 3, further comprising storing at leasta portion of the second amount of energy generated from the secondtorque at the energy storage facility.
 5. The method of claim 3, furthercomprising operating at least one of the one or more active suspensionactuators in at least three of four quadrants of a force/velocitydomain.
 6. The method of claim 3, further comprising updating the firsttorque at a frequency of at least 1 Hertz.
 7. The method of claim 3,further comprising operating at least one of the one or more activesuspension actuators in at all four quadrants of a force/velocitydomain.
 8. The method of claim 2, wherein the one or more bypass valvesincludes a blow-off valve.
 9. The method of claim 2, further comprisingcontrolling at least one of the one or more bypass valves to selectivelybypass the flow of hydraulic fluid.
 10. The method of claim 2, whereinmovement of the wheel comprises a wheel event.
 11. An active suspensionsystem for a vehicle, the active suspension system comprising: one ormore active suspension actuators, each actuator comprising an actuatorbody, a compression volume, an extension volume, and a piston; ahydraulic pump in fluid communication with the compression volume andthe extension volume; an electric motor operatively coupled to thehydraulic pump; an energy storage facility electrically connected to theelectric motor; a controller configured to send one or more torquecommands to the electric motor and configured to source an amount ofenergy from the energy storage facility; one or more sensors configuredto detect movement of a wheel of the vehicle and send data representingthe movement of the wheel to the controller; and one or more bypassvalves configured to modify flow of hydraulic fluid to the hydraulicpump from at least one of the compression volume or the extensionvolume, wherein the controller is configured to determine and send oneor more torque commands to the electric motor and configured to sourceenergy from the energy storage facility based at least partially on datareceived from at least one of the one or more sensors, and wherein thecontroller is configured to apply a first torque corresponding to afirst torque command to a shaft of the hydraulic pump using at least aportion of the required amount of energy, thereby altering a pressure inat least one of the compression volume or the extension volume.
 12. Theactive suspension system of claim 11, further comprising an accumulatorthat is in fluid communication with at least one of the compressionvolume and the extension volume.
 13. The active suspension system ofclaim 12, wherein the electric motor is configured to be driven as agenerator using a second torque produced at the hydraulic pump due toanother road induced movement of the wheel to generate a second amountof energy.
 14. The active suspension system of claim 13, wherein theenergy storage facility is configured to store at least a portion of thesecond amount of energy generated from the second torque.
 15. The activesuspension system of claim 13, wherein the active suspension system isconfigured to operate at least one of the one or more active suspensionactuators in at least three of four quadrants of a force/velocitydomain.
 16. The active suspension system of claim 13, wherein thecontrolleris configured to update the first torque at a frequency of atleast 1 Hertz.
 17. The active suspension system claim 13, wherein theactive suspension system is configured to operate at least one of theone or more active suspension actuators in at all four quadrants of aforce/velocity domain.
 18. The active suspension system of claim 12,wherein the one or more bypass valves includes a blow-off valve.
 19. Theactive suspension system of claim 12, wherein the controller isconfigured to control at least one of the one or more bypass valves toselectively bypass the flow of hydraulic fluid.
 20. The activesuspension system claim 12, wherein movement of the wheel comprises awheel event.